Continuously variable transmission

ABSTRACT

A variable speed transmission having a plurality of tilting balls and opposing input and output discs is illustrated and described that provides an infinite number of speed combinations over its transmission ratio range. The use of a planetary gear set allows minimum speeds to be in reverse and the unique geometry of the transmission allows all of the power paths to be coaxial, thereby reducing overall size and complexity of the transmission in comparison to transmissions achieving similar transmission ratio ranges.

RELATED APPLICATIONS

This application is a continuation of, and incorporates by reference inits entirety, U.S. patent application Ser. No. 10/788,736, filed on Feb.26, 2004, which claims priority from U.S. Provisional Application No.60/450,965 filed Feb. 28, 2003, U.S. Provisional Application No.60/494,376 filed Aug. 11, 2003, U.S. Provisional Application No.60/512,600 filed Oct. 16, 2003 and U.S. Provisional Application60/537,938 filed Jan. 21, 2004. The entire disclosure of each of theseapplications is hereby incorporated by reference.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The field of the invention relates generally to transmissions, and moreparticularly the invention relates to continuously variabletransmissions.

2. Description of the Related Art

In order to provide a continuously variable transmission, varioustraction roller transmissions in which power is transmitted throughtraction rollers supported in a housing between torque input and outputdiscs have been developed. In such transmissions, the traction rollersare mounted on support structures which, when pivoted, cause theengagement of traction rollers with the torque discs in circles ofvarying diameters depending on the desired transmission ratio.

However, the success of these traditional solutions has been limited.For example, in one solution, a driving hub for a vehicle with avariable adjustable transmission ratio is disclosed. This method teachesthe use of two iris plates, one on each side of the traction rollers, totilt the axis of rotation of each of the rollers. However, the use ofiris plates can be very complicated due to the large number of partsthat are required to adjust the iris plates during transmissionshifting. Another difficulty with this transmission is that it has aguide ring that is configured to be predominantly stationary in relationto each of the rollers. Since the guide ring is stationary, shifting theaxis of rotation of each of the traction rollers is difficult.

One improvement over this earlier design includes a shaft about which aninput disc and an output disc rotate. The input disc and output disc areboth mounted on the shaft and contact a plurality of balls disposedequidistantly and radially about the shaft. The balls are in frictionalcontact with both discs and transmit power from the input disc to theoutput disc. An idler located concentrically over the shaft and betweenthe balls applies a force to keep the balls separate so as to makefrictional contact against the input disc and output disc. A keylimitation of this design is the absence of means for generating andadequately controlling the axial force acting as normal contact force tokeep the input disc and output disc in sufficient frictional contactagainst the balls as the speed ratio of the transmission changes. Due tothe fact that rolling traction continuously variable transmissionsrequire more axial force at low speed to prevent the driving and drivenrotating members from slipping on the speed changing friction balls,excessive force is applied in high speed and at a 1:1 ratio, when theinput and output speeds are equal. This excessive axial force lowersefficiency and causes the transmission to fail significantly faster thanif the proper amount of force was applied for any particular gear ratio.The excessive force also makes it more difficult to shift thetransmission.

Therefore, there is a need for a continuously variable transmission withan improved axial load generating system that changes the force producedas a function of the transmission ratio.

SUMMARY OF THE INVENTION

The systems and methods illustrated and described herein have severalfeatures, no single one of which is solely responsible for its desirableattributes. Without limiting the scope as expressed by the descriptionthat follows, its more prominent features will now be discussed briefly.After considering this discussion, and particularly after reading thesection entitled “Detailed Description of the Preferred Embodiments” onewill understand how the features of the system and methods provideseveral advantages over traditional systems and methods.

In a first aspect, a variable speed transmission is disclosed,comprising a longitudinal axis, a plurality of balls distributedradially about the longitudinal axis, each ball having a tiltable axisabout which it rotates, a rotatable input disc positioned adjacent tothe balls and in contact with each of the balls, a rotatable output discpositioned adjacent to the balls opposite the input disc and in contactwith each of the balls, a rotatable idler having a substantiallyconstant outer diameter coaxial about the longitudinal axis andpositioned radially inward of and in contact with each of the balls, anda planetary gear set mounted coaxially about the longitudinal axis ofthe transmission.

Embodiments of the variable speed transmission are also disclosedwherein the balls sum a torque component transmitted from at least twopower paths, which power paths are provided by the planetary gear setand wherein the at least two power paths are coaxial. In anotherembodiment, the at least one of the idler and the output disc provide atorque input to the planetary gearset.

In another aspect, a variable speed transmission is disclosed whereinthe planetary gearset further comprises; a ring gear mounted coaxiallyabout the longitudinal axis and having teeth that face radially inwardtowards, a plurality of planet gears distributed coaxially about thelongitudinal axis within the ring gear and in engagement with the ringgear, each planet gear having a respective planet axis about which itrotates, and wherein the planet axes are located radially away from thelongitudinal axis, a plurality of planet shafts, one for each planet,about which the planet gears rotate, a sun gear mounted coaxially aboutthe longitudinal axis and radially within and in engagement with each ofthe plurality of planet gears, and a planet carrier mounted coaxiallyabout the longitudinal axis and adapted to support and position theplanet shafts.

Some of these embodiments further comprise a cage adapted to align thetiltable axes of the balls and further adapted to maintain the angularand radial positions of the balls. In some embodiments, an input torqueis supplied to the planet carrier and the planet carrier is coupled tothe input disc, wherein the sun gear is coupled to the cage, wherein thering gear is fixed and does not rotate, and wherein an output torque issupplied from the transmission by the output disc.

In another aspect an axial force generator is disclosed for use withtransmission embodiments described herein that is adapted to generate anaxial force that increases the traction between the input disc, theballs, the idler and the output disc. In some embodiments, an amount ofaxial force generated by the axial force generator is a function of thetransmission ratio of the transmission. In other embodiments, each ofthe input disc, the balls, the output disc, and the idler have contactsurfaces that are coated with a friction increasing coating material.The coating material of certain embodiments is a ceramic or a cermet. Inyet other embodiments, the coating is a material selected from the groupconsisting of silicon nitride, silicon carbide, electroless nickel,electroplated nickel, or any combination thereof.

In yet another aspect, a variable speed transmission is disclosedcomprising; a longitudinal axis, a plurality of balls distributedradially about the longitudinal axis, each ball having a tiltable axisabout which it rotates, a rotatable input disc positioned adjacent tothe balls and in contact with each of the balls, a fixed output discpositioned adjacent to the balls opposite the input disc and in contactwith each of the balls, a rotatable idler having a constant outsidediameter and positioned radially inward of and in contact with each ofthe balls, a cage, adapted to maintain the radial position and axialalignment of the balls and that is rotatable about the longitudinalaxis, and an idler shaft connected to the idler adapted to receive atorque output from the idler and transmit the torque output out of thetransmission.

In still another aspect, a variable speed transmission is describedcomprising; first and second pluralities of balls distributed radiallyabout the longitudinal axis, first and second rotatable input discs, aninput shaft coaxial with the longitudinal axis and connected to thefirst and second input discs, a rotatable output disc positioned betweenthe first and second pluralities of balls and in contact with each ofthe first and second pluralities of balls, a first generally cylindricalidler positioned radially inward of and in contact with each of thefirst plurality of balls, and a second generally cylindrical idlerpositioned radially inward of and in contact with each of the secondplurality of balls.

For use with many embodiments described herein there is also disclosedan axial force generator adapted to apply an axial force to increasecontact force between the input disc, the output disc and the pluralityof speed adjusters, the axial force generator further comprising, abearing disc coaxial with and rotatable about the longitudinal axishaving an outer diameter and an inner diameter and having a threadedbore formed in its inner diameter, a plurality of perimeter rampsattached to a first side of the bearing disc near its outer diameter, aplurality of bearings adapted to engage the plurality of bearing discramps, a plurality of input disc perimeter ramps mounted on the inputdisc on a side opposite of the speed adjusters adapted to engage thebearings, a generally cylindrical screw coaxial with and rotatable aboutthe longitudinal axis and having male threads formed along its outersurface, which male threads are adapted to engage the threaded bore ofthe bearing disc, a plurality of central screw ramps attached to thescrew, and a plurality of central input disc ramps affixed to the inputdisc and adapted to engage the plurality of central screw ramps.

In another aspect, a support cage is disclosed that supports andpositions a plurality of speed adjusting tiltable balls in a rollingtraction transmission, which utilizes an input disc and an output discon either side of the plurality of balls, the cage comprising; first andsecond flat support discs that are each a generally circular sheethaving a plurality of slots extending radially inward from an outeredge, each slot having two sides, and a plurality of flat supportingspacers extending between said first and second support discs eachspacer having a front side, a back side, a first end and a second end,wherein the first and second ends each have a mounting surface, whereineach mounting surface has a curved surface, and wherein the spacers arepositioned angularly about the support discs between the grooves in thesupport discs such that the curved surfaces are aligned with the sidesof the grooves.

In yet another aspect, a support leg for a ratio changing mechanism,which changes the transmission ratio in a rolling traction transmissionby tilting an axle that forms the axis of rotation of aratio-determining ball, is disclosed that comprises; an elongated body,an axle-connecting end, a cam end opposite the axle-connecting end, afront side that faces the ball and a backside that faces away from theball, and a central support portion between the axle-connecting end andthe cam end, wherein the axle-connecting end has a bore formed throughit adapted to receive the axle, and wherein a convexly curved cammingsurface is formed on the front side of the cam end that is adapted toassist in controlling the alignment of the bore.

Another aspect is disclosed for a fluid pumping ball for use in avariable speed rolling traction transmission utilizing a plurality ofballs rotatable about their respective tiltable axes, an input disc onone side of and in contact with each of the plurality of balls, and anoutput disc on another side of and in contact with each of the pluralityof balls, the fluid pumping ball comprising; a spherical ball having abore formed through a diameter of the ball creating a cylindrical innersurface through the ball, and at least one helical groove formed in theinner surface of the ball and extending through the ball.

In still another aspect a fluid pumping axle is disclosed for use in avariable speed rolling traction transmission utilizing a plurality ofballs having respective axes formed by diametrical bores formedtherethrough, an input disc on one side of and in contact with each ofthe plurality of balls, and an output disc on another side of and incontact with each of the plurality of balls, the fluid pumping axlecomprising a generally cylindrical axle of a diameter smaller than thatof the bore through the balls and having first and second ends and amiddle region, wherein when the axle is positioned properly within thebore of its respective ball, the first and second ends extend out ofopposite sides of the ball and the middle region resides within theball, and at least one helical groove formed on an outside surface ofthe axle, wherein the helical groove begins at a point outside of theball and extends into at least a portion of the middle region.

In another embodiment, a shifting mechanism is disclosed for a variablespeed rolling traction transmission having a longitudinal axis and thatutilizes a plurality of tilting balls distributed in planar alignmentabout the longitudinal axis and each ball contacted on opposing sides byan input disc and an output disc, in order to control a transmissionratio of the transmission, the shifting mechanism comprising a tubulartransmission axle running along the longitudinal axis, a plurality ofball axles each extending through a bore formed through a correspondingone of the plurality of balls and forming a tiltable axis of thecorresponding ball about which that ball spins, and each ball axlehaving two ends that each extend out of the ball, a plurality of legs,one leg connected to each of the ends the ball axles, the legs extendingradially inward toward the transmission axle, an idler having asubstantially constant outside diameter that is positioned coaxiallyabout the transmission axle and radially inward of and in contact witheach of the balls, two disc-shaped shift guides, one on each end of theidler, and each having a flat side facing the idler and a convex curvedside facing away from the idler, wherein shift guides extend radially tocontact all of the respective legs on the corresponding side of theballs, a plurality of roller pulleys, one for each leg, wherein eachroller pulley is attached to a side of its respective leg facing awayfrom the balls, a generally cylindrical pulley stand extending axiallyfrom at least one of the shift guides, a plurality of guide pulleys, onefor each roller pulley, distributed radially about and attached to thepulley stand, and a flexible tether having first and second ends withthe first end extending through the axle and out a slot, which is formedin the axle proximate to the pulley stand, the first end of the tetherfurther wrapping around each of the roller pulleys and each of the guidepulleys, wherein the second end extends out of the axle to a shifter,wherein the guide pulleys are each mounted upon one or more pivot jointsto maintain alignment of each guide pulley with its respective rollerpulley and wherein when the tether is pulled by the shifter, the secondend draws each of the roller pulleys in to shift the transmission.

In another embodiment, a shifting mechanism is disclosed for a variablespeed transmission having a longitudinal axis and that utilizes aplurality of tilting balls, each having a ball radius from respectiveball centers, in order to control a transmission ratio of thetransmission, comprising a plurality of ball axles each extendingthrough a bore formed through a corresponding ball and forming thetiltable axis of the corresponding ball, and each ball axle having twoends that each extend out of the ball, a plurality of legs, one legconnected to each of ends the ball axles, the legs extending radiallyinward toward the transmission axle, a generally cylindrical idler witha substantially constant radius positioned coaxially and radially inwardof and in contact with each of the balls, first and second disc-shapedshift guides, one on each end of the idler, and each having a flat sidefacing the idler and a convex curved side facing away from the idler,wherein shift guides extend radially to contact all of the respectivelegs on the corresponding side of the balls, and a plurality of guidewheels each having a guide wheel radius, one guide wheel for each leg,each guide wheel rotatably mounted at a radially inward end of itsrespective leg, wherein the guide wheels contact the curved surface ofits respective shift guide, wherein a shapes of the convex curves aredetermined by a set of two-dimensional coordinates, the origin of iscentered at the intersection of the longitudinal axis and a line drawnthrough the centers of any two diametrically opposing balls, wherein thecoordinates represent the location of the point of contact between theguide wheel surface and the shift guide surface as a function of theaxial movement of the idler and shift guide, assuming that the convexcurve is substantially tangent to the guide wheel at the point ofcontact.

In still another embodiment, an automobile is disclosed, comprising anengine, a drivetrain; and a variable speed transmission comprising alongitudinal axis, a plurality of balls distributed radially about thelongitudinal axis, each ball having a tiltable axis about which itrotates, a rotatable input disc positioned adjacent to the balls and incontact with each of the balls, a rotatable output disc positionedadjacent to the balls opposite the input disc and in contact with eachof the balls, a rotatable idler having a substantially constant outerdiameter coaxial about the longitudinal axis and positioned radiallyinward of and in contact with each of the balls, and a planetary gearset mounted coaxially about the longitudinal axis of the transmission.

These and other improvements will become apparent to those skilled inthe art as they read the following detailed description and view theenclosed figures.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a cutaway side view of an embodiment of the transmissionshifted into high.

FIG. 2 is a cutaway side view of the transmission of FIG. 1 shifted intolow.

FIG. 3 is a partial end cross-sectional view of the transmission takenon line III-III of FIG. 1.

FIG. 4 is a schematic cutaway side view of the idler and rampsub-assembly of the transmission of FIG. 1.

FIG. 5 is a schematic perspective view of the ball sub-assembly of thetransmission of FIG. 1.

FIG. 6 is a schematic view of the shift rod sub-assembly of thetransmission of FIG. 1.

FIG. 7 is a schematic cutaway side view of the cage sub-assembly of thetransmission of FIG. 1.

FIG. 8 is a cutaway side view of the output disc of the transmission ofFIG. 1.

FIG. 9 is a schematic cutaway perspective view of the transmission ofFIG. 1.

FIG. 10 is a schematic cutaway side view of an alternative embodiment ofthe axial force generator of the transmission of FIG. 1.

FIG. 11 is a cutaway side view of an alternative embodiment of thetransmission of FIG. 1.

FIG. 12 is a schematic cutaway of the cage sub-assembly of thetransmission of FIG. 11.

FIG. 13 is a cutaway schematic view of an alternative disengagementmechanism looking from near the axis of the transmission of FIG. 11.

FIG. 14 is a cutaway schematic view of an alternative disengagementmechanism looking from above and outside the transmission of FIG. 11toward the center.

FIG. 15 is a cutaway schematic view of a portion of the axial forcegenerator sub-assembly of the transmission of FIG. 11.

FIG. 16 is a cutaway side view of the variator of the transmission ofFIG. 1.

FIG. 17 is a schematic cutaway side view of an alternative embodiment ofthe transmission of FIG. 1 with two variators.

FIG. 18 is a partial end cross-sectional view of the transmission takenon line I-I of FIG. 17.

FIG. 19 is a perspective view of the transmission of FIG. 17.

FIG. 20 is a perspective view of the iris plate of the transmission ofFIG. 17.

FIG. 21 is a perspective view of a stator of the transmission of FIG.17.

FIG. 22 is a cutaway side view of an alternate cage of the transmissionof FIG. 17.

FIG. 23 is a cutaway side view of a ball with grooves of the ball/legassembly of FIG. 5.

FIG. 24 is a cutaway side view of an alternate leg of the ball/legassembly of FIG. 5.

FIG. 25 is a schematic illustration of the ball and leg assembly showingapplicable geometric relations used to create a convex curves for theshift guides of the transmissions of FIGS. 1 and 17.

FIG. 26 is a schematic illustration of the ball and leg assembly in atilted orientation showing applicable geometric relations used to createthe convex curves for the shift guides of the transmissions of FIGS. 1and 17.

FIG. 27 is a schematic illustration of the convex curves illustratingcertain geometric relations utilized to create a convex curve for theshift guides of the transmissions of FIGS. 1 and 17.

FIG. 28 is a schematic view of the transmission of FIG. 1 showing itsfunction as a planetary gearset.

FIG. 29 is a schematic view of the transmission of FIG. 1 showing thethree planet gears in a first ratio.

FIG. 30 is a schematic view of the transmission of FIG. 1 showing thethree planet gears in a second ratio.

FIG. 31 is a schematic view of the transmission of FIG. 1 showing thethree planet gears in a third ratio.

FIG. 32 is a schematic view of the transmission of FIG. 1 combined witha planetary gearset on the output side and a parallel power path.

FIG. 33 is a schematic view of the transmission of FIG. 1 combined witha planetary gearset on the input side and a parallel power path.

FIG. 34 is a schematic view of the transmission of FIG. 1 combined witha planetary gearset on the output side.

FIG. 35 is a schematic perspective view of the transmission of FIG. 1combined with a planetary gearset on the input side.

FIGS. 36 a, b, and c are a cross-sectional side view, a perspectiveendview, and a schematic skeleton diagram, respecitvely, of anembodiment of an infinitely variable transmission utilizing one torqueinput and providing two sources of torque output.

FIG. 37 a is a cross-sectional side view of an alternative embodiment ofa continuously variable transmission where the output disc is part of arotating hub.

FIG. 37 b is a cross-sectional side view of an alternative embodiment ofa continuously variable transmission where the output disc is part of astationary hub.

FIG. 38 is a side view of an alternative ball axle.

FIG. 39 a is a cross-sectional side view of alternative axial forcegenerator for any of the transmission embodiments described herein.

FIGS. 39 b and c are a cross-sectional view and a perspective view,respectively, of a screw of the alternative axial force generator.

FIG. 40 a is a side elevation view of an alternate linkage assembly foruse with the alternate axial force generator of FIG. 39.

FIG. 40 b is a side elevation view of the alternate linkage assembly ofFIG. 40 a in an extended configuration.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

Embodiments of the invention will now be described with reference to theaccompanying figures, wherein like numerals refer to like elementsthroughout. The terminology used in the description presented herein isnot intended to be interpreted in any limited or restrictive mannersimply because it is being utilized in conjunction with a detaileddescription of certain specific embodiments of the invention.Furthermore, embodiments of the invention may include several novelfeatures, no single one of which is solely responsible for its desirableattributes or which is essential to practicing the inventions hereindescribed.

The transmissions described herein are of the type that utilize speedadjuster balls with axes that tilt as described in U.S. Pat. Nos.6,241,636, 6,322,475, and 6,419,608. The embodiments described in thesepatents and those described herein typically have two sides generallyseparated by a variator portion, to be described below, an input sideand an output side. The driving side of the transmission, that is theside that receives the torque or the rotational force into thetransmission is termed the input side, and the driven side of thetransmission or the side that transfers the torque from the transmissionout of the transmission is termed the output side. An input disc and anoutput disc are in contact with the speed adjuster balls. As the ballstilt on their axes, the point of rolling contact on one disc movestoward the pole or axis of the ball, where it contacts the ball at acircle of decreasing diameter, and the point of rolling contact on theother disc moves toward the equator of the ball, thus contacting thedisc at a circle of increasing diameter. If the axis of the ball istilted in the opposite direction, the input and output discsrespectively experience the converse relationship. In this manner, theratio of rotational speed of the input disc to that of the output disc,or the transmission ratio, can be changed over a wide range by simplytilting the axes of the speed adjuster balls. The centers of the ballsdefine the border between the input side and the output side of thetransmission and similar components that are located on both the inputside of the balls and the output side of the balls are generallydescribed herein with the same reference numbers. Similar componentslocated on both the input and output sides of the transmission generallyhave the suffix “a” attached at the end of the reference number if theyare located on the input side, and the components located on the outputside of the transmission generally have the suffix “b” attached at theend of their respective reference numbers.

Referring to FIG. 1, an embodiment of a transmission 100 is illustratedhaving a longitudinal axis 11 about which multiple speed adjusting balls1 are radially distributed. The speed adjusting balls 1 of someembodiments stay in their angular positions about the longitudinal axis11, while in other embodiments the balls 1 are free to orbit about thelongitudinal axis 11. The balls 1 are contacted on their input side byan input disc 34 and on their output side by an output disc 101. Theinput and out put discs 34, 101 are annular discs extending from aninner bore near the longitudinal axis on their respective input andoutput sides of the balls 1 to a radial point at which they each makecontact with the balls 1. The input and output discs 34, 101 each have acontact surface that forms the contact area between each disc 34 and101, and the balls 1. In general, as the input disc 34 rotates about thelongitudinal axis 11, each portion of the contact area of the input disc34 rotates and sequentially contacts each of the balls 1 during eachrotation. This is similar for the output disc 101 as well. The inputdisc 34 and the output disc 101 can be shaped as simple discs or can beconcave, convex, cylindrical or any other shape, depending on theconfiguration of the input and output desired. In one embodiment theinput and output discs are spoked to make them lighter for weightsensitive applications. The rolling contact surfaces of the discs wherethey engage the speed adjuster balls can have a flat, concave, convex orother shaped profile, depending on the torque and efficiencyrequirements of the application. A concave profile where the discscontact the balls decreases the amount of axial force required toprevent slippage while a convex profile increases efficiency.Additionally, the balls 1 all contact an idler 18 on their respectiveradially innermost point. The idler 18 is a generally cylindricalcomponent that rests coaxially about the longitudinal axis 11 andassists in maintaining the radial position of the balls 1. Withreference to the longitudinal axis 11 of many embodiments of thetransmission, the contact surfaces of the input disc 34 and the outputdisc 101 can be located generally radially outward from the center ofthe balls 1, with the idler 18 located radially inward from the balls 1,so that each ball 1 makes three-point contact with the idler 18, theinput disc 34, and the output disc 101. The input disc 34, the outputdisc 101, and the idler 18 can all rotate about the same longitudinalaxis 11 in many embodiments, and are described in fuller detail below.

Due to the fact that the embodiments of transmissions 100 describedherein are rolling traction transmissions, in some embodiments, highaxial forces are required to prevent slippage of the input disc 34 andoutput disc 101 at the ball 1 contacts. As axial force increases duringperiods of high torque transfer, deformation of the contact patcheswhere the input disc 34, the output disc 101, and the idler 18 contactthe balls 1 becomes a significant problem, reducing efficiency and thelife of these components. The amount of torque that can be transferredthrough these contact patches is finite and is is a function of theyield strength of the material from which the balls 1, the input disc,34, the output disc 101, and the idler 18 are made. The frictioncoefficient of the balls 1, the input disc, 34, the output disc 101, andthe idler 18 has a dramatic effect on the amount of axial force requiredto transfer a given amount of torque and thus greatly affects theefficiency and life of the transmission. The friction coefficient of therolling elements in a traction transmission is a very important variableaffecting performance.

Certain coatings may be applied to the surfaces of the balls 1, theinput disc, 34, the output disc 101, and the idler 18 to improve theirperformance. In fact, such coatings can be used advantageously on therolling contacting elements of any rolling traction transmission toachieve the same added benefits that are achieved for the embodiments oftransmissions described herein. Some coatings have the beneficial effectof increasing the friction coefficient of the surfaces of these rollingelements. Some coatings have a high friction coefficient and alsodisplay a variable coefficient of friction, which increases as axialforce increases. A high friction coefficient allows less axial force tobe required for a given torque, thereby increasing efficiency and lifeof the transmission. A variable coefficient of friction increases themaximum torque rating of the transmission by decreasing the amount ofaxial force required to transfer this maximum torque.

Some coatings, such as ceramics and cermets, possess excellent hardnessand wear properties, and can greatly extend the life of the highlyloaded rolling elements in a rolling traction transmission. A ceramiccoating such as silicon nitride can have a high friction coefficient, avariable coefficient of friction which increases as axial forceincreases, and can also increase the life of the balls 1, the inputdisc, 34, the output disc 101, and the idler 18 when applied to thesurfaces of these components in a very thin layer. The coating thicknessdepends on the material used for the coating and can vary fromapplication to application but typically is in the range of 0.5 micronsto 2 microns for a ceramic and 0.75 microns to 4 microns for a cermet.

The process used to apply the coating is important to consider when theballs 1, the input disc, 34, the output disc 101, and the idler 18 aremade from hardened steel, which is the material used in many embodimentsof the transmissions described herein. Some processes used to applyceramics and cermets require high temperatures and will lower thehardness of the balls 1, the input disc, 34, the output disc 101, andthe idler 18, harming performance and contributing to premature failure.A low temperature application process is desirable and several areavailable, including low temperature vacuum plasma, DC pulsed reactivemagnetron sputtering, plasma- enhanced chemical vapor deposition(PE-CVD), unbalanced magnetron physical vapor deposition, and plating.The plating process is attractive due to its low cost and because acustom bath can be created to achieve desired coating properties.Immersing the rolling elements in a bath of silicon carbide or siliconnitride with co-deposited electroless nickel or electroplated nickelwith silicon carbide or silicon nitride is a low temperature solutionthat is well suited for high volume production. It should be noted thatother materials can be used in addition to those mentioned. With thisapplication process, the parts are contained in a cage, immersed in thebath, and shaken so that the solution contacts all surfaces. Thicknessof the coating is controlled by the length of time that the componentsare immersed in the bath. For instance, some embodiments will soak thecomponents using silicon nitride with co-deposited electroless nickelfor four (4) hours to achieve the proper coating thickness, althoughthis is just an example and many ways to form the coating and controlits thickness are known and can be used taking into account the desiredproperties, the desired thickness and the substrate or base metal ofwhich the components are made.

FIGS. 1, 2, and 3 illustrate an embodiment of a continuously variabletransmission 100 that is shrouded in a case 40 which protects thetransmission 100, contains lubricant, aligns components of thetransmission 100, and absorbs forces of the transmission 100. A case cap67 can, in certain embodiments, cover the case 40. The case cap 67 isgenerally shaped as a disc with a bore, through its center through whichan input shaft passes, and that has a set of threads at its outerdiameter that thread into a corresponding set of threads on the innerdiameter of the case 40. Although in other embodiments, the case cap 67can be fastened to the case 40 or held in place by a snap ring andcorresponding groove in the case 40, and would therefore not need to bethreaded at its outer diameter. In embodiments utilizing fasteners toattach the case cap 67, the case cap 67 extends to the inside diameterof the case 40 so that case fasteners (not shown) used to bolt the case40 to the machinery to which the transmission 100 is attached can bepassed through corresponding holes in the case cap 67. The case cap 67of the illustrated embodiment has a cylindrical portion extending froman area near its outer diameter toward the output side of thetransmission 100 for additional support of other components of thetransmission 100. At the heart of the illustrated transmission 100embodiment is a plurality of balls 1 that are typically spherical inshape and are radially distributed substantially evenly or symmetricallyabout the centerline, or longitudinal axis 11 of rotation of thetransmission 100. In the illustrated embodiment, eight balls 1 are used.However, it should be noted that more or fewer balls 1 could be useddepending on the use of the transmission 100. For example, thetransmission may include 3, 4, 5, 6, 7, 8, 9, 10, 11, 12, 13, 14, 15 ormore balls. The provision for more than 3, 4, or 5 balls can more widelydistribute the forces exerted on the individual balls 1 and their pointsof contact with other components of the transmission 100 and can alsoreduce the force necessary to prevent the transmission 100 from slippingat the ball 1 contact patches. Certain embodiments in applications withlow torque but a high transmission ratio use few balls 1 of relativelylarger diameters, while certain embodiments in applications with hightorque and a high transmission ratio can use more balls 1 or relativelylarger diameters. Other embodiments, in applications with high torqueand a low transmission ratio and where high efficiency is not important,use more balls 1 of relatively smaller diameters. Finally, certainembodiments, in applications with low torque and where high efficiencyis not important, use few balls 1 of relatively smaller diameters.

Ball axles 3 are inserted through holes that run through the center ofeach of the balls 1 to define an axis of rotation for each of the balls1. The ball axles 3 are generally elongated shafts over which the balls1 rotate, and have two ends that extend out of either side of the holethrough the balls 1. Certain embodiments have cylindrically shaped ballaxles 3, although any shape can be used. The balls 1 are mounted tofreely rotate about the ball axles 3.

In certain embodiments, bearings (not separately illustrated) areutilized to reduce the friction between the outer surface of the ballaxles 3 and the surface of the bore through the corresponding ball 1.These bearings can be any type of bearings situated anywhere along thecontacting surfaces of the balls 1 and their corresponding ball axles 3,and many embodiments will maximize the life and utility of such bearingsthrough standard mechanical principles common in the design of dynamicmechanical systems. In some of these embodiments, radial bearings arelocated at each end of the bore through the balls 1. These bearings canincorporate the inner surface of the bore or the outer surface of theball axles 3 as their races, or the bearings can include separate racesthat fit in appropriate cavities formed in the bore of each ball 1 andon each ball axle 3. In one embodiment, a cavity (not shown) for abearing is formed by expanding the bore through each ball 1 at least atboth ends an appropriate diameter such that a radial bearing, roller,ball or other type, can be fitted into and held within the cavity thusformed. In another embodiment, the ball axles 3 are coated with afriction reducing material such as babbit, Teflon or other suchmaterial.

Many embodiments also minimize the friction between the ball axles 3 andthe balls 1 by introducing lubrication in the bore of the ball axles 3.The lubrication can be injected into the bore around the ball axles 3 bya pressure source, or it can be drawn into the bore by the rifling orhelical grooves formed on the ball axles 3 themselves. Furtherdiscussion of the lubrication of the ball axles 3 is provided below.

In FIG. 1, the axes of rotation of the balls 1 are shown tilted in adirection that puts the transmission in a high ratio, wherein the outputspeed is greater than the input speed. If the ball axles 3 arehorizontal, that is parallel to the main axis of the transmission 100,the transmission 100 is in a 1:1 input rotation rate to output rotationrate ratio, wherein the input and output rotation speeds are equal. InFIG. 2, the axes of rotation of the balls 1 are shown tilted in adirection where the transmission 100 is in a low ratio, meaning theoutput rotation speed is slower than the input rotation speed. For thepurpose of simplicity, only the parts that change position ororientation when the transmission 100 is shifted are numbered in FIG. 2.

FIGS. 1, 2, 4, and 5 illustrate how the axes of the balls 1 can betilted in operation to shift the transmission 100. Referring to FIG. 5,a plurality of legs 2, which in most embodiments are generally struts,are attached to the ball axles 3 near each of the ends of the ball axles3 that extend beyond the ends of the holes bored through the balls 1.Each leg 2 extends from its point of attachment to its respective ballaxle 3 radially inward toward the axis of the transmission 100. In oneembodiment, each of the legs 2 has a through bore that receives arespective end of one of the ball axles 3. The ball axles 3 preferablyextend through the legs 2 such that they have an end exposed beyond eachleg 2. In the illustrated embodiments, the ball axles 3 advantageouslyhave rollers 4 coaxially and slidingly positioned over the exposed endsof the ball axles 3. The rollers 4 are generally cylindrical wheelsfitted over the ball axles 3 outside of and beyond the legs 2 and rotatefreely about the ball axles 3. The rollers 4 can be attached to the ballaxles 3 via spring clips or other such mechanism, or they can ridefreely over the ball axles 3. The rollers 4 can be radial bearings forinstance, where the outer races of the bearings form the wheel orrolling surface. As illustrated in FIGS. 1 and 7, the rollers 4 and theends of the ball axles 3 fit inside grooves 86 formed by or in a pair ofstators 80 a, 80 b.

The stators 80 a, 80 b of one embodiment are illustrated in FIGS. 5 and7. The illustrated input stator 80 a and output stator 80 b aregenerally in the form of parallel discs annularly located about thelongitudinal axis 11 of the transmission on either side of the balls 1.The stators 80 a, 80 b of many embodiments are comprised of input statordiscs 81 a and output stator discs 81 b, respectively, which aregenerally annular discs of substantially uniform thickness with multipleapertures to be discussed further below. Each input and output statordisc 81 a, 81 b has a first side that faces the balls 1 and a secondside that faces away from the balls 1. Multiple stator curves 82 areattached to the first side of the stator discs 81 a, 81 b. The statorcurves 82 are curved surfaces attached or affixed to the stator discs 81a, 81 b that each have a concave face 90 facing toward the balls 1 and aconvex face 91 facing away from the balls 1 and contacting theirrespective stator discs 81. In some embodiments, the stator curves 82are integral with the stator discs 81 a, 81 b. The stator curves 82 ofmany embodiments have a substantially uniform thickness and have atleast one aperture (not separately shown) used to align and attach thestator curves 82 to each other and to the stator discs 81. The statorcurves 82 of many embodiments, or the stator discs 81 a, 81 b whereintegral parts are used, include a slot 710 that accepts a flat spacer83, which allows further positioning and alignment of the stator curves82 and stator discs 81 a, 81 b. The flat spacers 83 are generally flatand generally rectangular pieces of rigid material that extend betweenand interconnect the input stator 80 a and the output stator 80 b. Theflat spacers 83 fit within the slots 710 formed in the stator curves 82.In the illustrated embodiment, the flat spacers 83 are not fastened orotherwise connected to the stator curves 82, however, in someembodiments the flat spacers 83 are attached to the stator curves 82 bywelding, adhesive, or fastening.

Also illustrated in FIG. 7, multiple cylindrical spacers 84, of agenerally cylindrical shape with bores at least in each end, areradially positioned inside of the flat spacers 83 and also connect andposition the stator discs 81 and stator curves 82. The bores of thecylindrical spacers 84 accept one spacer fastener 85 at each end. Thespacer fasteners 85 are designed to clamp and hold the stator discs 81a, 81 b, the stator curves 82, the flat spacers 83, and the cylindricalspacers 84 together, which collectively form the cage 89. The cage 89maintains the radial and angular positions of the balls 1 and aligns theballs 1 with respect to one another.

The rotational axes of the balls 1 are changed by moving either theinput-side or output-side legs 2 radially out from the axis of thetransmission 100, which tilts the ball axles 3. As this occurs, eachroller 4 fits into and follows a groove 86, which is slightly largerthan the diameter of the roller 4, and is formed by the space betweeneach pair of adjacent stator curves 82. The rollers 4 therefore rollalong the surface of the sides 92, 93 of the stator curves 82, a firstside 92 and a second side 93 for each stator curve 82, in order tomaintain the plane of movement of the ball axles 3 in line with thelongitudinal axis 11 of the transmission 100. In many embodiments, eachroller 4 rolls on a first side 92 of the stator curve 82 on the inputside of the transmission 100 and on the corresponding first side 92 ofthe corresponding output stator curve 82. Typically in such embodiments,the forces of the transmission 100 prevent the rollers 4 from contactingthe second side 93 of the stator curves 82 in normal operation. Therollers 4 are slightly smaller in diameter than the width of the grooves86 formed between the stator curves 82, forming a small gap between theedges of the grooves 86 and the circumference of each correspondingroller. If the opposing sets of stator curves 82 on the input stator 80a and output stator 80 b were in perfect alignment, the small gapbetween the circumferences of the rollers 4 and the grooves 86 wouldallow the ball axles to slightly tilt and become misaligned with thelongitudinal axis 11 of the transmission 100. This condition producessideslip, a situation where the balls axles 3 are allowed to slightlymove laterally, which lowers overall transmission efficiency. In someembodiments, the stator curves 82 on the input and output sides of thetransmission 100 may be slightly offset from each other so that the ballaxles 3 remain parallel with the axis of the transmission 100. Anytangential force, mainly a transaxial force, the balls 1 may apply tothe ball axles 3 is absorbed by the ball axles 3, the rollers 4 and thefirst sides 92, 93 of the stator curves 82. As the transmission 100 isshifted to a lower or higher transmission ratio by changing therotational axes of the balls 1, each one of the pairs of rollers 4,located on the opposite ends of a single ball axle 3, move in oppositedirections along their respective corresponding grooves 86 by rolling upor down a respective side of the groove 86.

Referring to FIGS. 1 and 7, the cage 89 can be rigidly attached to thecase 40 with one or more case connectors 160. The case connectors 160extend generally perpendicularly from the radial outermost part of theflat spacers 83. The case connectors 160 can be fastened to the flatspacers 83 or can be formed integrally with the flat spacers 83. Theoutside diameter formed roughly by the outsides of the case connectors160 is substantially the same dimension as the inside diameter of thecase 40 and holes in both the case 40 and case connectors 160 providefor the use of standard or specialty fasteners, which rigidly attach thecase connectors 160 to the case 40, thus bracing and preventing the cage40 from moving. The case 40 has mounting holes providing for theattachment of the case 40 to a frame or other structural body. In otherembodiments, the case connectors 160 can be formed as part of the case40 and provide a location for attachment of the flat spacers 83 or othercage 89 component in order to mobilize the cage 89.

FIGS. 1, 5, and 7 illustrate an embodiment including a pair of statorwheels 30 attached to each of the legs 2 that roll on the concave face90 of the curved surfaces 82 along a path near the edge of the sides 92,93. The stator wheels 30 are attached to the legs 2 generally in thearea where the ball axles 3 pass through the legs 2. The stator wheels30 can be attached to the legs 2 with stator wheel pins 31, which passthrough a bore through the legs 2 that is generally perpendicular to theball axles 3, or by any other attachment method. The stator wheels 30are coaxially and slidingly mounted over the stator wheel pins 31 andsecured with standard fasteners, such as snap rings for example. In someembodiments, the stator wheels 30 are radial bearings with the innerrace mounted to the stator wheel pins 31 and the outer race forming therolling surface. In certain embodiments, one stator wheel 30 ispositioned on each side of a leg 2 with enough clearance from the leg 2to allow the stator wheels 30 to roll radially along the concave faces90, with respect to the longitudinal axis 11 of the transmission 100,when the transmission 100 is shifted. In certain embodiments, theconcave faces 90 are shaped such that they are concentric about a radiusfrom the longitudinal axis 11 of the transmission 100 formed by thecenter of the balls 1.

Still referring to FIGS. 1, 5, and 7, guide wheels 21 are illustratedthat can be attached to the end of the legs 2 that are nearest thelongitudinal axis 11 of the transmission 100. In the illustratedembodiment, the guide wheels 21 are inserted into a slot formed in theend of the legs 2. The guide wheels 21 are held in place in the slots ofthe legs 21 with guide wheel pins 22, or by any other attachment method.The guide wheels 21 are coaxially and slidingly mounted over the guidewheel pins 22, which are inserted into bores formed in the legs 2 oneach side of the guide wheels 21 and perpendicular to the plane of theslot. In some embodiments, the legs 2 are designed to elasticallydeflect relatively slightly in order to allow for manufacturingtolerances of the parts of the transmission 100. The ball 1, the legs 2,the ball axle 3, the rollers 4, the stator wheels 30, the stator wheelpins 31, the guide wheels 21, and the guide wheel pins 22 collectivelyform the ball/leg assembly 403 seen in FIG. 5.

Referring to the embodiment illustrated in FIGS. 4, 6, and 7, shiftingis actuated by rotating a rod 10 that is positioned outside of the case40. The rod 10 is utilized to wrap an unwrap a flexible input cable 155a and a flexible output cable 155 b that are attached to, at theirrespective first ends, and wrapped around the rod 10, in oppositerespective directions. In some embodiments, the input cable 155 a iswrapped counter-clockwise around the rod 10 and the output cable 155 bis wrapped clockwise around the rod 10, when looking from right to leftas the rod 10 is illustrated in FIG. 6. Both the input cable 155 a andthe output cable 155 b extend through holes in the case 40 and thenthrough the first end of an input flexible cable housing 151 a, and anoutput flexible cable housing 151 b. The input flexible cable housing151 a and the output flexible cable housing 151 b of the illustratedembodiment are flexible elongated tubes that guide the input cable 155 aand output cable 155 b radially inward toward the longitudinal axis 11then longitudinally out through holes in the stator discs 81 a, b andthen again radially inward where the second end of the input and outputflexible cable housings 151 a, b are inserted into and attach to thefirst end of input and output rigid cable housings 153 a, b,respectively. The input and output rigid cable housings 153 a, b, areinflexible tubes through which the cables 155 a, b, pass and are guidedradially inward from the second ends of the flexible cable housings 151a, b and then direct the cables 155 a, b longitudinally through holes inthe stator discs 81 a, b and toward a second end of the rigid cablehousings 153 a, b near the idler 18. In many embodiments, the cables 155a, b are attached at their second ends to an input shift guide 13 a, andan output shift guide 13 b (described further below) with conventionalcable fasteners, or other suitable attachment means. As will bediscussed further below, the shift guides 13 a, 13 b position the idler18 axially along the longitudinal axis 11 and position the legs 3radially, thereby changing the axes of the balls 1 and the ratio of thetransmission 100.

If the rod 10 is rotated counter-clockwise, relative to the axis of therod 10 from right to left as illustrated in FIG. 6, by the user, eithermanually or by or assisted with a power source, the input cable 155 aunwinds from the rod 10 and the output cable 155 b winds onto the rod10. Therefore, the second end of the output cable 155 b applies atension force to the output shift guide 13 b and the input cable 155 ais unwinding a commensurate amount from the rod 10. This moves the idler18 axially toward the output side of the transmission 100 and shifts thetransmission 100 toward low.

Still referring to FIGS. 4, 5, and 7, the illustrated shift guides 13 a,b, are each generally of the form of an annular ring with inside andoutside diameters, and are shaped so as to have two sides. The firstside is a generally straight surface that dynamically contacts andaxially supports the idler 18 via two sets of idler bearings 17 a, 17 b,which are each associated with a respective shift guide 13 a, b. Thesecond side of each shift guide 13 a, b, the side facing away from theidler 18, is a cam side that transitions from a straight or flat radialsurface 14, towards the inner diameter of the shift guides 13 a, b, to aconvex curve 97 towards the outer diameter of the shift guides 13 a, b.At the inner diameter of the shift guides 13 a, b a longitudinal tubularsleeve 417 a, b extends axially toward the opposing shift guide 13 a, bin order to mate with the tubular sleeve 417 a, b from that shift guide13 a, b. In some embodiments, as illustrated in FIG. 4, the tubularsleeve of the input side shift guide 13 a has part of its inner diameterbored out to accept the tubular sleeve of the output shift guide 13 b.Correspondingly, a portion of the outer diameter of the tubular sleeveof the output shift guide 13 b has been removed to allow a portion ofthat tubular sleeve 417 a, b to be inserted into the tubular sleeve 417a, b of the input shift guide 13 a. This provides additional stabilityto the shift guides 13 a, b of such embodiments.

The cross section side view of the shift guides 13 a, b illustrated inFIG. 4 shows that, in this embodiment, the flat surface 14 profile ofthe side facing away from the is perpendicular to the longitudinal axis11 up to a radial point where the guide wheels 21 contact the shiftguides 13 a, b, if the ball axles 3 are parallel with the longitudinalaxis 11 of the transmission 100. From this point moving out toward theperimeter of the shift guide 13 a, b the profile of the shift guides 13a, b curves in a convex shape. In some embodiments, the convex curve 97of a shift guide 13 a, b is not a radius but is composed of multipleradii, or is shaped hyperbolically, asymptotically or otherwise. As thetransmission 100 is shifted toward low, the input guide wheels 21 a,roll toward the longitudinal axis 11 on the flat 14 portion of shiftguide 13 a, and the output guide wheels 21 b roll on the convex curved97 portion of the shift guide 13 b away from the longitudinal axis 11.The shift guides 13 a, b, can be attached to each other by eitherthreading the tubular sleeve of the input shift guide 13 a with malethreads and the tubular sleeve of the output sleeve 13 b with femalethreads, or vice versa, and threading the shift guides 13 a, b,together. One shift guide 13 a, b, either the input or output, can alsobe pressed into the other shift guide 13 a, b. The shift guides 13 a, bcan also be attached by other methods such as glue, metal adhesive,welding or any other means.

The convex curves 97 of the two shift guides 13 a, b, act as camsurfaces, each contacting and pushing the multiple guide wheels 21. Theflat surface 14 and convex curve 97 of each shift guide 13 a, b contactthe guide wheels 21 so that as the shift guides 13 a, b, move axiallyalong the longitudinal axis 11, the guide wheels 21 ride along the shiftguide 13 a, b surface 14, 97 in a generally radial direction forcing theleg 2 radially out from, or in toward, the longitudinal axis 11, therebychanging the angle of the ball axle 3 and the rotational axis of theassociated ball 1.

Referring to FIGS. 4 and 7, the idler 18 of some embodiments is locatedin a trough formed between the first sides and the sleeve portions ofthe shift guides 13 a, b, and thus moves in unison with the shift guides13 a, b. In certain embodiments, the idler 18 is generally tubular andof one outside diameter and is substantially cylindrical along thecentral portion of its inside diameter with an input and output idlerbearing 17 a, b, on each end of its inside diameter. In otherembodiments, the outer diameter and inside diameters of the idler 18 canbe non-uniform and can vary or be any shape, such as ramped or curved.The idler 18 has two sides, one near the input stator 80 a, and one nearthe output stator 80 b. The idler bearings 17 a, 17 b provide rollingcontact between the idler 18 and the shift guides 13 a, b. The idlerbearings 17 a, 17 b are located coaxially around the sleeve portion ofthe shift guides 13 a, b, allowing the idler 18 to freely rotate aboutthe axis of the transmission 100. A sleeve 19 is fit around thelongitudinal axis 11 of the transmission 100 and fitting inside theinside diameters of the shift guides 13 a, b. The sleeve 19 is agenerally tubular component that is held in operable contact with aninside bearing race surface of each of the shift guides 13 a, b by aninput sleeve bearing 172 a and an output sleeve bearing 172 b. Thesleeve bearings 172 a, b, provide for rotation of the sleeve 19 byrolling along an outer bearing race complimentary to the races of theshift guides 13 a, b. The idler 18, the idler bearings 17 a, 17 b, thesleeve 19, the shift guides 13 a, 13 b, and the sleeve bearings 172 a,172 b collectively form the idler assembly 402, seen in FIG. 4.

Referring to FIGS. 4, 7, and 8, the sleeve 19 of some embodiments hasits inside diameter threaded to accept the threaded insertion of anidler rod 171. The idler rod 171 is a generally cylindrical rod thatlies along the longitudinal axis 11 of the transmission 100. In someembodiments, the idler rod 171 is threaded at least partially along itslength to allow insertion into the sleeve 19. The first end of the idlerrod 171, which faces the output side of the transmission 100, ispreferably threaded through the sleeve 19 and extends out past theoutput side of the sleeve 19 where it is inserted into the insidediameter of the output disc 101.

As illustrated in FIG. 8, the output disc 101 in some embodiments isgenerally a conical disc that is spoked to reduce weight and has atubular sleeve portion extending from its inner diameter axially towardthe output side of the transmission 100. The output disc 101 transfersthe output torque to a drive shaft, wheel, or other mechanical device.The output disc 101 contacts the balls 1 on their output side androtates at a speed different than the input rotation of the transmissionat ratios other than 1:1. The output disc 101 serves to guide and centerthe idler rod 171 at its first end so that the sleeve 19, idler 18, andshift guides 13 a, b stay concentric with the axis of the transmission100. Alternately, an annular bearing may be positioned over the idlerrod 171, between the idler rod 171 and the inside diameter of the outputdisc 101, to minimize friction. The idler rod 171, sleeve 19, shiftguides 13 a, b, and idler 18 are operably connected, and all moveaxially in unison when the transmission 100 is shifted.

Referring to FIG. 2, a conical spring 133, positioned between the inputshift guide 13 a and stator 80 a biases the shifting of the transmission100 toward low. Referring to FIG. 1, output disc bearings 102, whichcontact a bearing race near the perimeter of the output disc 101, absorband transfer axial force generated by the transmission 100 to the case40. The case 40 has a corresponding bearing race to guide the outputdisc bearings 102.

Referring to FIGS. 4, 5, and 7, the limits of the axial movement of theshift guides 13 a, b define the shifting range of the transmission 100.Axial movement is limited by inside faces 88 a, b, on the stator discs81 a, b, which the shift guides 13 a, b, contact. At an extreme hightransmission ratio, shift guide 13 a contacts the inside face 88 a onthe input stator discs 81 a, and at an extreme low transmission ratio,the shift guide 13 b contacts the inside face 88 on the output statordisc 81 b. In many embodiments, the curvature of the convex curves 97 ofthe shift guides 13 a, b, is functionally dependent on the distance fromthe center of a ball 1 to the center of the guide wheel 21, the radiusof the guide wheel 21, the angle between lines formed between the twoguide wheels 21 and the center of the ball 1, and the angle of tilt ofthe ball 1 axis. An example of such a relationship is described below,with respect to FIGS. 25, 26 and 27.

Now referring to embodiments illustrated by FIGS. 1, 5, and 7, one ormore stator wheels 30 can be attached to each leg 2 with a stator wheelpin 31 that is inserted through a hole in each leg 2. The stator wheelpins 31 are of the proper size and design to allow the stator wheels 30to rotate freely over each stator wheel pin 31. The stator wheels 30roll along the concave curved surfaces 90 of the stator curves 82 thatface the balls 1. The stator wheels 30 provide axial support to preventthe legs 2 from moving axially and also to ensure that the ball axles 3tilt easily when the transmission 100 is shifted.

Referring to FIGS. 1 and 7, a spoked input disc 34, located adjacent tothe stator 80 a, partially encapsulates but generally does not contactthe stator 80 a. The input disc 34 may have two or more spokes or may bea solid disc. The spokes reduce weight and aid in assembly of thetransmission 100. In other embodiments a solid disc can be used. Theinput disc 34 has two sides, a first side that contacts with the balls1, and a second side that faces opposite the first side. The input disc34 is generally an annular disk that fits coaxially over, and extendsradially from, a set of female threads or nut 37 at its inner diameter.The outside diameter of the input disc 34 is designed to fit within thecase 40, if the case 40 used is the type that encapsulates the balls 1and the input disc 34 and mounts to a rigid support structure 116 suchas a chassis or frame with conventional bolts, which are insertedthrough bolt holes in a flange on the case 40. As mentioned above, theinput disc 34 is in rotating contact with the balls 1 along acircumferential ramped or bearing contact surface on a lip of the firstside of the input disc 34, the side facing the balls 1. As alsomentioned above, some embodiments of the input disc 34 have a set offemale threads 37, or a nut 37, inserted into its inside diameter, andthe nut 37 is threaded over a screw 35, thereby engaging the input disc34 with the screw 35.

Referring to FIGS. 1 and 4, the screw 35 is attached to and rotated by adrive shaft 69. The drive shaft 69 is generally cylindrical and has aninner bore, a first end facing axially towards the output side, a secondend facing axially toward the input side, and a generally constantdiameter. At the first end, the drive shaft 69 is rigidly attached toand rotated by the input torque device, usually a gear, a sprocket, or acrankshaft from a motor. The drive shaft 69 has axial splines 109extending from its second end to engage and rotate a corresponding setof splines formed on the inside diameter of the screw 35. A set ofcentral drive shaft ramps 99, which on a first side is generally a setof raised inclined surfaces on an annular disk that is positionedcoaxially over the drive shaft 69, have mating prongs that mate with thesplines 109 on the drive shaft 99, are rotated by the drive shaft 69,and are capable of moving axially along the drive shaft 69. A pin ring195 contacts a second side of the central drive shaft ramps 99. The pinring 195 is a rigid ring that is coaxially positioned over the idler rod171, is capable of axial movement and has a transverse bore thatfunctions to hold an idler pin 196 in alignment with the idler rod 171.The idler pin 196 is an elongated rigid rod that is slightly longer thanthe diameter of the pin ring 195 and which is inserted through anelongated slot 173 in the idler rod 171 and extends slightly beyond thepin ring 195 at both its first and second ends when it is inserted intothe bore of the pin ring 195. The elongated slot 173 in the idler rod171 allows for axial movement of the idler rod 171 to the right, whenviewed as illustrated in FIG. 1, without contacting the pin 196 when thetransmission 100 is shifted from 1:1 toward high. However, when thetransmission 100 is shifted from 1:1 toward low, the side on the inputend of the elongated slot 173 contacts the pin 196, which then operablycontacts the central drive shaft ramps 99 via the pin ring 195. Theidler rod 171 is thus operably connected to the central drive shaftramps 99 when the transmission is between 1:1 and low so that when theidler rod 171 moves axially the central drive shaft ramps 99 also moveaxially in conjunction with the idler rod 171. The ramp surfaces of thecentral drive shaft ramps 99 can be helical, curved, linear, or anyother shape, and are in operable contact with a set of correspondingcentral bearing disc ramps 98. The central bearing disc ramps 98 haveramp faces that are complimentary to and oppose the central drive shaftramps 99. On a first side, facing the output side of the transmission100, the central bearing disc ramps 98 face the central drive shaftramps 99 and are contacted and driven by the central drive shaft ramps99.

The central bearing disc ramps 98 are rigidly attached to a bearing disc60, a generally annular disc positioned to rotate coaxially about thelongitudinal axis 11 of the transmission 100. The bearing disc 60 has abearing race near its perimeter on its side that faces away from theballs 1 that contacts a bearing disc bearing 66. The bearing discbearing 66 is an annular thrust bearing at the perimeter of the bearingdisc 60 and is positioned between the bearing disc 60 and the input disc34. The bearing disc bearing 66 provides axial and radial support forthe bearing disc 60 and in turn is supported by a bearing race on a casecap 67, which acts with the case 40 to partially encapsulate the innerparts of the transmission 100.

Referring to FIG. 1, the case cap 67 is generally an annular discextending from the drive shaft 69 having a tubular portion extendingtoward the output end from at or near its perimeter and also having abore through its center. The case cap 67 absorbs axial and radial forcesproduced by the transmission 100, and seals the transmission 100,thereby preventing lubricant from escaping and contamination fromentering. The case cap 67 is stationary and, in some embodiments, isrigidly attached to the case 40 with conventional fastening methods orcan have male threads on its outside diameter, which mate withcorresponding female threads on the inside diameter of the case 40. Aswas mentioned above, the case cap 67 has a bearing race that contactsthe bearing disc bearing 66 near the perimeter of the bearing disc 60that is located at the inside of the output end of the tubular extensionfrom the case cap 67. The case cap 67 also has a second bearing racefacing the output side located near the inside diameter of its annularportion that mates with a drive shaft bearing 104. The drive shaftbearing 104 is a combination thrust and radial bearing that providesaxial and radial support to the drive shaft 69. The drive shaft 67 has abearing race formed on its outside diameter facing the input side thatmates with the drive shaft bearing 104, which transfers the axial forceproduced by the screw 35 to the case cap 67. An input bearing 105, addssupport to the drive shaft 69. The input bearing 105 is coaxiallypositioned over the drive shaft 69 and mates with a third race on theinside diameter of the case cap 67 facing the input side of thetransmission 100. A cone nut 106, a generally cylindrical threaded nutwith a bearing race designed to provide a running surface for the inputbearing 105, is threaded over the drive shaft 69 and supports the inputbearing 105.

Referring to the embodiment illustrated in FIG. 1, a set of multipleperimeter ramps 61, generally forming a ring about the longitudinal axis11, are rigidly attached to the bearing disc 60. The perimeter ramps 61are multiple inclined surfaces that are positioned radially about thelongitudinal axis 11 and are positioned against or formed on the bearingdisc 60 and face the output side. The inclined surfaces can be curved,helical, linear, or another shape and each one creates a wedge thatproduces and axial force that is applied to one of multiple rampbearings 62. The ramp bearings 62 are spherical but can be cylindrical,conical, or another geometric shape, and are housed in a bearing cage63. The bearing cage 63 of the illustrated embodiment is generally ringshaped with multiple apertures that contain the individual ramp bearings62. A set of input disc ramps 64 are rigidly attached to, or formed aspart of, the input disc 34. The input disc ramps 64 in some embodimentsare complimentary to the perimeter ramps 62 with the ramps facing towardthe input side. In another embodiment, the input disc ramps 64 are inthe form of a bearing race that aligns and centers the ramp bearings 62radially. The ramp bearings 62 respond to variations in torque byrolling up or down the inclined faces of the perimeter ramps 61 and theinput disc ramps 64.

Referring now to FIGS. 1 and 4, an axial force generator 160 is made upof various components that create an axial force that is generated andis applied to the input disc 34 to increase the normal contact forcebetween the input disc 34 and the balls 1, which is a component in thefriction the input disc 34 utilizes in rotating the balls 1. Thetransmission 100 produces sufficient axial force so that the input disc34, the balls 1, and the output disc 101 do not slip, or slip only anacceptable amount, at their contact points. As the magnitude of torqueapplied to the transmission 100 increases, an appropriate amount ofadditional axial force is required to prevent slippage. Furthermore,more axial force is required to prevent slippage in low than in high orat a 1:1 speed ratio. However, providing too much force in high or at1:1 will shorten the lifespan of the transmission 100, reduceefficiency, and/or necessitate larger components to absorb the increasedaxial forces. Ideally, the axial force generator 160 will vary the axialforce applied to the balls 1 as the transmission 100 is shifted and alsoas torque is varied. In some embodiments, the transmission 100accomplishes both these goals. The screw 35 is designed and configuredto provide an axial force that is separate and distinct from thatproduced by the perimeter ramps 61. In some embodiments the screw 35produces less axial force than the perimeter ramps 61, although in otherversions of the transmission 100, the screw 35 is configured to producemore force than the perimeter ramps 61. Upon an increase in torque, thescrew 35 rotates slightly farther into the nut 37 to increase axialforce by an amount proportional to the increase in torque. If thetransmission 100 is in a 1:1 ratio and the user or vehicle shifts into alower speed, the idler rod 171, moves axially toward the input side,along with the sleeve 19, sleeve bearings 172, shift guides 13 a, b, andidler 18. The idler rod 171 contacts the central drive shaft ramps 99through the pin 196 and pin ring 195, causing the central drive shaftramps 99 to move axially toward the output side. The ramped surfaces ofthe central drive shaft ramps 99 contact the opposing ramped surfaces ofthe central bearing disc ramps 98, causing the central bearing discramps 98 to rotate the bearing disc 67 and engage the perimeter ramps 61with the ramp bearings 62 and the input disc ramps 64. The central driveshaft ramps 99 and the central bearing disc ramps 98 perform a torquesplitting function, shifting some of the torque from the screw 35 to theperimeter ramps 61. This increases the percentage of transmitted torquethat is directed through the perimeter ramps 61, and due to the fact theperimeter ramps 61 are torque sensitive as described above, the amountof axial force that is generated increases.

Still referring to FIGS. 1 and 4, when shifting into low, the idler 18moves axially towards the output side, and is pulled toward low by areaction of forces in the contact patch. The farther the idler 18 movestoward low, the stronger it is pulled. This “idler pull,” whichincreases with an increase in normal force across the contact as well asshift angle, also occurs when shifting into high. The idler pull occursdue to a collection of transverse forces acting in the contact patch,the effect of which is called spin. Spin occurs at the three contactpatches, the points of contact where the balls contact the input disc34, the output disc 101, and the idler 18. The magnitude of theresultant forces from spin at the contact between the idler 18 and theballs 1 is minimal in comparison to that of the balls 1 and input andoutput discs 34, 101. Due to the minimal spin produced at the contactpatch of the idler 18 and ball 1 interface, this contact patch will beignored for the following explanation. Spin can be considered anefficiency loss in the contact patches at the input disc 34 and ball 1and also at the output disc 101 and ball 1. Spin produces a transverseforce perpendicular to the rolling direction of the balls 1 and discs34, 101. At a 1:1 ratio the transverse forces produced by spin, orcontact spin, at the input and output contact patches are equal andopposite and are essentially cancelled. There is no axial pull on theidler 18 in this condition. However, as the transmission 100 is shiftedtoward low for example, the contact patch at the input disc 34 and ball1 moves farther from the axis or pole of the ball 1. This decreases spinas well as the transverse forces that are produced perpendicular to therolling direction. Simultaneously the output disc 101 and ball 1 contactpatch moves closer to the axis or pole of the ball 1, which increasesspin and the resultant transverse force. This creates a situation wherethe transverse forces produced by spin on the input and output sides ofthe transmission 100 are not equal and because the transverse force onthe output contact is greater, the contact patch between the output disc101 and ball 1 moves closer to the axis of the ball 1. The farther thetransmission 100 is shifted into low the stronger the transverse forcesin the contacts become that are exerted on the ball 1. The transverseforces caused by spin on the ball 1 exert a force in the oppositedirection when shifting into high. The legs 2 attached to the ball axles3 transfer the pull to the shift guides 13 a, b, and because the shiftguides 13 a, b, are operably attached to the idler 18 and sleeve 19, anaxial force is transferred to the idler rod 171. As the normal forceacross the contact increases, the influence of contact spin increases atall ratios and efficiency decreases.

Still referring to FIGS. 1 and 4, as the transmission 100 is shiftedinto low, the pull transferred to the idler rod 171 results in an axialforce toward the left, as viewed in FIG. 1, which causes the inputtorque to shift from the screw 35 to the perimeter ramps 61. As thetransmission 100 is shifted into extreme low, the idler rod 171 pullsmore strongly, causing relative movement between the central drive shaftramps 99 and the central bearing disc ramps 98 and shifts even moretorque to the perimeter ramps 61. This reduces the torque transmittedthrough the screw 35 and increases the torque transmitted through theperimeter ramps 61, resulting in an increase in axial force.

Referring to FIGS. 1 and 9, a disengagement mechanism (composed ofseveral parts to be described) is described. The disengagement mechanismis located between the input disc 35 and the bearing disc 60 anddisengages the transmission 100 when output rotation is greater thaninput rotation. The disengagement mechanism is comprised of multipleparts, including an input disc connector 121, a generally cylindricalelongated pin that is rigidly attached to the input disc 34 near itsperimeter, which protrudes from the input disc 35 towards the bearingdisc 60 in a direction substantially parallel to the longitudinal axis11 of the transmission 100. The input disc connector 121 engages aclutch lever 122 at a first end. The clutch lever 122 is a generallyL-shaped flat piece of rigid material, having its first end extending asits short leg and a second end extending as its long leg, and thatpivots on a preloader 123 from a joint at the intersection of its legs.The engagement of the input disc connector 121 and the first end of theclutch lever 122 is sliding engagement and allows relative movementbetween the input disc connector 121 and the clutch lever 122. Theclutch lever 122 joint is formed by a through hole that is positionedover the preloader 123. The preloader 123 is a flexible, elongated rodthat can also be square, flat, or of any other cross-sectional shape andis attached at one of its ends to a hole extending radially through thebearing cage 63, and at a second end is rigidly attached to the driveshaft 69. The preloader 123 can bias the ramp bearings 62 up theperimeter ramps 61, it can pull the input disc 34 off of the ballsduring times when the disengagement mechanism is activated, and it canserve as a means of attachment for other components, such asdisengagement mechanism 120 components. A pawl 124 is also attached tothe clutch lever 122. The pawl 124 is generally wedge-shaped and at afirst end tapers to a point, and at a second end is rounded with athrough hole. A pawl pin 125 is inserted into a hole in the second endof the clutch lever 122, thereby attaching the pawl 124 to the clutchlever 122 while allowing for rotation of the pawl 124 about the pawl pin125. The pawl 124 mates with and contacts a disc shaped ratchet 126,which has teeth around its circumference and lays flat against the backof the clutch lever 122. At the center of the ratchet 126 is a holethrough which the preloader 123 passes adjacent to the clutch lever 122and radially inward toward the longitudinal axis 11 of the transmission100. The ratchet 126 is held in place by conventional fasteners and iscapable of rotation about the preloader 123. A ratchet bevel 127, a gearwith beveled teeth around its perimeter, is rigidly and coaxiallyattached to and made a part of the ratchet 126. The teeth on the ratchetbevel 127 mesh with a bevel gear 128. The bevel gear 128 is a ring thatis rigidly attached to the bearing disc 60 in the illustratedembodiment, but which can be attached to other rotating components suchas the drive shaft 69 and central drive shaft ramps 99. The bevel gear128 has teeth around its perimeter that mate with the teeth on theratchet 126. A main spring 129, a coiled spring with multiple coils asillustrated in FIG. 11, is coaxially positioned around the longitudinalaxis 11 of the transmission 100 and is attached at a first end to theinput disc 34 and at a second end to the bearing disc 60. The mainspring 129 biases the input disc 34 to rotate about or “unwind” from thescrew 35 so that the input disc 34 contacts the balls 1.

Still referring to FIGS. 1 and 9, when input rotation to thetransmission 100 ceases and the output disc 101 continues to be rotatedby one or more wheels, a drive train, or other output rotatingmechanism, the balls 1 are driven by the output disc 101. The balls 1then rotate the input disc 34 in a first direction to “wind” onto thescrew 35 and disengage from the balls 1. The input disc connector 121,rotated by the input disc 34 in the same first direction, contacts androtates the clutch lever 122 and also the pawl 124 in a first direction.The pawl 124 is biased to contact the teeth of the ratchet 126 by a pawltensioner (not shown), which can be a torsion spring positionedcoaxially over the pawl pin 125. As the pawl 124 passes over the teethof the ratchet 126, the pawl 124 locks onto the teeth of the ratchet126, preventing the input disc 34 from unwinding off of the screw 35 ina second direction and again contacting the balls 1, as the bias of themain spring 129 would tend to do. The ratchet 126 is prevented fromrotating in a second direction because the ratchet bevel 127, a part ofthe ratchet 126, has teeth that are interlocked with the bevel gear 128which is not rotating.

When input rotation of the transmission 100 resumes, the bevel gear 127is rotated by the bearing disc 60 in a first direction, which rotatesthe ratchet bevel 127 and ratchet 126 in a second direction, thusrotating the pawl 124 in a second direction, allowing the main spring129 to bias the input disc 34 to unwind from the screw 35 in a seconddirection and contact the balls 1. It is important to note that thebearing cage 63, attached to the preloader 123 at a first end, causesthe preloader 123 to rotate relative to the input disc 34 when the inputdisc 34 rotates in a first direction. This is due to the ramp bearings62 rotating relative to the input disc 34 when the input disc 34 isrotating in a first direction. Similarly, when input rotation of thetransmission 100 resumes, the bearing disc 60 rotates relative to thepreloader 123 due to the same relative rotation. This action providesfor the engagement and release of the disengagement mechanism 120.

Referring to FIGS. 1 and 15, a latch 115 rigidly attaches to the side ofthe input disc 34 that faces the bearing disk 60 and engages a hook 114that is rigidly attached to a first of two ends of a hook lever 113. Thehook lever 113 is an elongated strut with the hook 114 at its first endand a hook hinge 116 at its second end. The latch 115 has an engagingarea or an opening that is larger than the width of the hook 114 andprovides extra room for the hook 114 to move radially, with respect tothe longitudinal axis 11, within the confines of the latch 114 when theinput disc 34 and the bearing disk 60 move relative to each other. Thehood hinge 116 engages a middle hinge 119 and forms a hinge joint with afirst hinge pin 111. The middle hinge 119 is integral with a first endof an input disc lever 112, which is a generally elongated strut havingtwo ends. On its second end, the input disc lever 112 has an input dischinge 117, which engages a hinge brace 110 via the use of a second hingepin 118. The hinge brace 110 is generally a base to support the hook114, the hook lever 113, the hook hinge 116, the first hinge pin 111,the middle hinge 119, the input disc lever 112, the second hinge pin118, and the input disc hinge 117, and it is rigidly attached to thebearing disc 60 on the side facing the input disc 34. When the latch 115and hook 114 are engaged, the ramp bearings 62 are prevented fromrolling to an area on the perimeter ramps 61 that does not provide thecorrect amount of axial force to the drive disk 34. This positiveengagement ensures that all rotational force applied to the rampbearings 62 by perimeter ramps 61 is transmitted to the input disc 34. Apreloader 123 is attached at a first end to the drive shaft 69 andextends radially outward. At a second end the preloader contacts theinput disc lever 112, biasing the input disc 34 away from the balls 1,so that on occasions when the input disc 34 disengages from the balls 1,it is biased to remain disconnected.

Referring to FIG. 10, a cutaway side view of an alternative axial forcegenerator of the transmission 100 is disclosed. For purposes ofsimplicity, only the differences between the axial force generatorpreviously described and the axial force generator illustrated in FIG.10 will be presented. The illustrated axial force generator includes oneor more reversing levers 261. The reversing levers 261 are generallyflat, irregularly shaped cam pieces each having an off-center mountedpivot hole with a first side radially inward of the pivot hole and asecond side radially outside of the pivot hole. The first side of thereversing levers 261 each fit into the elongated slot .173 in the idlerrod 171. When the transmission 200 is shifted toward low, the end of theelongated slot 173 contacts the first side of the reversing levers 261and the reversing levers 261 pivot on an axis produced by a reversingpin 262 that is inserted into the pivot holes of the reversing levers261. As the first sides are contacted by the end of the elongated slot173, the first side of each of the reversing levers 261 moves toward theoutput side of the transmission 100 and the second side of the reversinglevers 261 moves toward the input side of the transmission 100 therebyfulfilling the cam function of the reversing levers 261. By increasingand decreasing the length of the first side and second side, thereversing levers 261 can be designed to decrease the distance that theymove axially toward the input side and increase the force they produce.The reversing levers 261 can be designed in this manner to create amechanical advantage to adjust the axial force that they produce. Attheir second sides, the reversing levers 261 each contact the outputside of the central screw ramps 298 when the transmission 100 is shiftedtoward low. The reversing levers 261 are each attached to a lever ring263 by the reversing pins 262, which can be pressed or threaded intoholes in the lever ring 263 to hold the reversing levers 261 inposition. The lever ring 263 is a ring shaped device that fits around,and slides axially along, the idler rod 171 and has one or morerectangular slots cut through it to allow for insertion and positioningof the reversing levers 261.

Still referring to the embodiment illustrated in FIG. 10, a set ofcentral screw ramps 299 is rigidly attached to and can be rotated by thescrew 35. The central screw ramps 299 of this embodiment are similar tothe central screw ramps 99 illustrated in FIG. 4, in that the centralscrew ramps 299 are formed as ramps on the second side of a disc havinga first side facing the output side and a second side facing the inputside. As the transmission 100 is shifted toward low, the second side ofthe reversing levers 261 pushes against the first side of the centralscrew ramps 299. The central screw ramps 299, which are splined to thedrive shaft 69 via the above-described spline 109, are rotated by thedrive shaft 69, are capable of axial movement along the longitudinalaxis 11, and are similar to the central drive shaft ramps 99 of theprevious embodiment, except that the central screw ramps 299 face theinput side of the transmission 100 rather than the output side. Thecentral screw ramps 299 contact an opposing set of central bearing discramps 298, which are free to rotate relative to the drive shaft 69 andare similar to the central bearing disc ramps 98 illustrated in FIG. 4,except that the central bearing disc ramps 298 face the output side ofthe transmission 100 rather than the input side. As the central screwramps 299 are pushed axially by the reversing levers 261 toward thecentral bearing disc ramps 298, relative rotation of the ramp faces ofthe central screw ramps 299 and central bearing disc ramps 298 isdeveloped that causes the bearing disc 60 to rotate to a point such thatthe perimeter ramps 61 become engaged, thereby shifting torque to theperimeter ramps 61 and increasing the amount of axial force that isgenerated.

Referring now to FIG. 11, a cutaway side view of an alternativeembodiment of the transmission 100 of FIG. 1 is disclosed. For purposesof simplicity, only the differences between the earlier transmission 100and this transmission 300 will be described. The transmission 300 has analternative cage 389, an alternative disengagement mechanism (item 320of FIGS. 13 and 14), and an alternative axial force generator.Furthermore, in the embodiment illustrated in FIG. 11 the conical spring133 is moved to the output side of the transmission 300, biasing theshifting toward high.

Referring now to FIGS. 11 and 12, an alternative cage 389 is disclosed.The cage 389 includes input and output stator discs 381 a, b, howeverfor ease of viewing, the output stator disc 381 b has been removed. Theoutput stator 381 b of many embodiments is structurally similar to theinput stator 381 a. Multiple stator curves 382 are attached to thestator discs 381 a, b and have first sides facing the balls 1 and secondsides facing away from the balls 1. The second side 391 of each of thestator curves 382 is a flat surface that lays flat against a respectiveone of the stator discs 181 a, b. The stator curves 382 have two throughholes that are used to attach the stator curves 382 to the stator discs381 a, b with conventional fasteners or other type of attachmentmechanism. The stator curves 382 have on each of their first sides arectangular slot into which multiple flat spacers 383 are inserted toconnect the stators 381. The flat spacers 383 serve to set the distancebetween the stators 381, create a strong connection between the stators381, and ensure that the stators 381 are parallel and in alignment.

The illustrated design incorporates a stator disc 181 that issubstantially flat. Therefore, the stator discs 181 can be manufacturedutilizing a substantially flat sheet of rigid material. The stator discs181 can be produced from any of a number of inexpensive manufacturingtechniques such as stamping, fine blanking, or any other such techniqueknown in the industry. The stator discs 181 of this design can be madefrom thin or sheet metal, plastic, ceramic, wood or paper products orany other material. The illustrated design allows for significantreduction in the cost of materials and manufacturing of these otherwiserelatively expensive components to a suitably high tolerance.

Referring now to FIGS. 11, 13, and 14, an alternative disengagementmechanism 320 is disclosed. FIG. 13 is a cutaway schematic view lookingfrom near the axis of the transmission 300, and FIG. 14 is a cutawayschematic view looking from above and outside the transmission 300generally radially inward toward the center. The ratchet 126 and theratchet bevel 127 of the previously described embodiment are merged inthe present embodiment into one pawl gear 326 that engages the pawl 124and has teeth that interlock with the bevel gear 328. The bevel gear 328in other embodiments may have non-beveled gear teeth. The clutch lever322 is a rigid, flat L-shaped component having three or more holes. Thecentermost hole at the joint of the two legs forming the “L” shapepositions the clutch lever 322 rotatably and coaxially about thepreloader 123. A hole near the end of the long leg of the clutch lever322 allows for insertion of the pawl pin 125 and attachment to the pawl124. A hole near the end of the short leg of the clutch lever 322 thatmates with the input disc connector 321, accepts and retains a clutchpin 329 which fits into a slot of the input disc connector 321. Theinput disc connector 321 is rigidly attached to the input disc 34 andhas a slot providing for sliding engagement of the clutch pin 329. Theoperation of the alternative disengagement mechanism 320 is otherwisethe same as the coasting mechanism 120 previously described andillustrated in FIGS. 1 and 9.

Referring now to FIGS. 11 and 15, an alternative axial force generatorincludes a generally conical wedge 360 that is positioned and is capableof axial movement along the central axis of the transmission 300. Theconical wedge 360 is also mated with the spline 109. As the transmission300 is shifted toward low, the conical wedge 360 is engaged by the idlerrod 171 and moves axially in the same direction as the idler rod 171.The conical wedge 360 contacts a first end of an AFG (axial forcegenerator) lever 362 near the transmission 300 axis. The AFG lever 362is a generally elongated part having a first semi-circular end thatengages the conical wedge 360 and then extends radially outward from thelongitudinal axis 11 to a second end that engages the input disc lever112. The AFG lever 362 is attached to the spline 109 with a fulcrum pin361 about which the AFG lever 362 rotates. The fulcrum pin 361 providesfor pivoting of the AFG lever 362 so that the second end of the AFGlever 362 engages the input disc lever 112. The input disc lever 112 isoperably attached to the bearing disc 60 and rotates the bearing disc 60so that the perimeter ramps 61 engage, thus shifting input torque fromthe screw 35 to the perimeter ramps 61. The operation of the alternativeaxial force generator 360 is otherwise the same as the axial forcegenerator previously described and seen in FIGS. 1 and 4.

Referring now to FIGS. 16 and 17, an alternative embodiment of thetransmission 100 of FIG. 1 is disclosed. For the purposes of simplicity,only those differences between the transmission 1700 of FIG. 17 and thetransmission 100 of FIG. 1 will be explained. The transmission 100 ofFIG. 1 includes one variator, The term variator can be used to describethe components of the transmission 100 that vary the input to outputspeed ratio. The assemblies and components comprising the variator 401of the present embodiment include the ball/leg assembly 403 of FIG. 5,the input disc 34, the output disc 101, the idler assembly 402 of FIG.4, and the cage 89 of FIG. 7. It should be noted that all components andassemblies of the variator 401 can change to best fit the specificapplication of the transmission 1700, and in FIG. 16 generic forms ofthe assemblies and components comprising the variator 401 are depicted.

The embodiment of the transmission 1700 illustrated in FIG. 17 issimilar to the transmission 100 of but includes two variators 401. Thisconfiguration is beneficial for applications where high torque capacityis required in a transmission 1700 with a small diameter or overallsize. This configuration also eliminates radial bearings needed tosupport the bearing disc 114 and the output disc 101, thereby increasingoverall efficiency. Due to the fact that the transmission 1700 has twovariators 401, each variator 401 has an output side and the transmission1700 also has an output side. Thus there are three output sides and inthis configuration the convention or marking like components with an “a”and a “b” to differentiate between the input and output sides is notused. However, as illustrated in FIG. 17, the input side is to the rightand the input is to the left.

Referring to FIGS. 17-19, a case 423 is illustrated that surrounds andencapsulates the transmission 1700. The case 423 is generallycylindrical and protects the transmission 1700 from outside elements andcontamination and additionally contains lubrication for properoperation. The case 423 is attached to an engine, frame, or other rigidbody (not shown) with standard fasteners (not shown), which fit throughcase holes 424. The case 423 is open on the input side, the side withthe case holes 424 or to the right as illustrated, to accept an inputtorque. Input torque is transmitted from an outside source to an inputshaft 425, which is a long, rigid, rod or shaft capable of transmittingtorque. The input shaft 425 transmits torque to a bearing disc 428 viasplines, keying, or other such manner. The bearing disc 428 is adisc-shaped rigid component capable of absorbing significant axialforces produced by the transmission 1700 and is similar in design to thebearing disc 60 illustrated in FIG. 1. An input shaft bearing 426 ispositioned coaxially over the input shaft 425 between a flange 429 onthe input end of the input shaft 425 and the bearing disc 428 to allow asmall amount of relative movement between the bearing disc 428 and theinput shaft 425. When the bearing disc 429 begins rotating, theperimeter ramps 61, ramp bearings 62, bearing cage 63, input disc ramps64, and input disc 34 rotate as previously described. This rotates theballs 1 in the first variator 420, is the one on the input side.

Simultaneously, as the input shaft 425 rotates a second input disc 431is rotated. The second input disc 431 is rigidly attached to the inputshaft 425, and can be keyed with a backing nut, pressed over the inputshaft 425, welded, pinned, or attached by other methods. The secondinput disc 431 is located on the output side of the transmission 1700,opposite the bearing disc 428. The second input disc 431 and the bearingdisc 428 absorb the considerable axial forces created by the perimeterramps 61, ramp bearings 62, and input disc ramps 64 that act as normalforces to prevent slippage at the ball/disc contact patches aspreviously described. The second input disc 431 is similar in shape tothe input disc 34 previously described and upon rotation of the inputshaft 425; it rotates the balls 1 in the second variator 422. The secondvariator 422 is generally a mirror image of the first variator 420 andis positioned farther from the input side of the transmission 1700 sothat the first variator 420 is situated between it and the input side.

As previously described, the balls 1 in the first variator 420 rotatethe output disc 430 through their rolling contact with that component.The output disc 430, although serving the same function as the outputdisc 101 previously described, has two opposing contact surfaces andcontacts balls 1 on both variators 420, 422. From the cross sectionalview illustrated in FIG. 17, the output disc 430 can be shaped in ashallow arch or upside down shallow “V,” the ends of which have acontact surface to contact the balls 1 of the two variators 420, 422.The output disc 430 surrounds the second variator 422 and extends towardthe output side in a generally cylindrical shape. In the illustratedembodiment, the cylindrical shape of the output disc 430 continuestoward the output side of the transmission 1700 surrounding the secondinput disc 431 after which the diameter of the output disc 430 decreasesand then again becomes a generally cylindrical shape of a smallerdiameter as it exits the case 423. To hold the output disc 430concentric and align it with the first and second input discs 34, 431,annular bearings 434, 435, may be used to radially align the output disc431. A case bearing 434 is positioned in the bore of the case 423 andover the output disc 430 and an output disc bearing 435 is positioned inthe bore of the output disc 430 and over the input shaft 425 to provideadditional support. The output disc 430 can be made of two pieces thatare connected together to form the illustrated output disc 430. Thisallows for assembly of the second variator 422 inside the cylindricalshell of the output disc 430.

As illustrated in FIG. 17, this can be accomplished by use of twoannular flanges along the large diameter of the output disc 430. In someembodiments, the annular flanges are located generally midway along thelarge diameter of the output disc 430. Referring now to FIGS. 17, 20,and 21, the ball axles 433 of the transmission 1700 are similar to theball axles 3 previously described and perform the same function. Inaddition, the ball axles 433 serve as the mechanism by which the balls 1are tilted to vary the speed ratio of the transmission 1700. The ballaxles 433 are elongated on each of their respective output sides andextend through the walls of the output stators 435. The output stators435 are similar to the output stators 80 b previously described, but themultiple radial grooves 436 penetrate all the way through the walls ofthe output stators 435. The grooves 436 of the output stators 435continue all the way through the output stator 435 walls so that aseries of equally spaced radial grooves 436 extend radially from nearthe bore at the center of the output stator 435 to the perimeter. Theball axles 433 have iris rollers 407 positioned coaxially over theirelongated output ends. The iris rollers 407 are generally cylindricalwheels that are capable of rotating over the ball axles 433 and aredesigned to fit inside the grooves 411 of an iris plate 409. The irisplate 409 is an annular disc or plate with a bore through its centerthat fits coaxially about the longitudinal axis 11 of the transmission1700. The iris plate 409 is of a thickness that is greater than twicethe thickness of each iris roller 407 and has a number of iris grooves411 extending radially outward from near the bore to near the perimeterof the iris plate 409. As the iris grooves 411 extend radially, theirangular position changes as well, so that as the iris plate 409 isrotated angularly about the longitudinal axis 11, the iris grooves 411provide a camming function along their respective lengths. In otherwords, the grooves 411 spiral out from near the bore in the center ofthe iris plate 409 to respective points near its perimeter.

The iris rollers 407 are radiused along their outside diameters, or havefillets on their outer corners, so that their diameters remain unchangedinside the grooves 411 of the iris plate 409 when the ball axles 433 aretilted. The iris plate 409 is of a thickness sufficient to allow irisrollers 407 from both variators 420, 422, to remain inside the grooves411 of the iris plate 433 at all shifting ratios. The iris grooves 411operate in traditional iris plate fashion and cause the ball axles 433to move radially inward or outward when the iris plate 409 is rotated.The iris plate 409 has a first side facing the first variator and asecond side facing the second variator and is coaxially positioned aboutthe longitudinal axis 11 of the transmission 1700 and over abuttingbosses on, tubular extensions extending from, the two output stators435. The two output stators 435 can be attached to each other withconventional fasteners through axial holes (not illustrated) in thebosses of the output stators 435. The output stator 435 bosses have ahole through their centers and multiple holes positioned radiallyoutward from the center. In some embodiments, the bosses on the outputstators 435 form a space slightly wider than the iris plate 409 toprovide freedom of rotation for the iris plate 433 and some embodimentsutilize bearings between the bosses and the iris plate 409 to accuratelycontrol the position of the iris plate 409 between the output stators435. An iris cable 406 is attached to the first side of the iris plate409 near the outside diameter of the iris plate 409 and extendslongitudinally from the point of connection. The iris cable 406 isrouted through the output stator 435 of the first variator 420 in anorientation so that when it is pulled, it rotates the iris plate 409.The iris cable 406, after passing through an aperture near the perimeterof the output stator 435 is routed through the case 423 to the outsideof the transmission 1700 where it allows for control of the transmissionratio. An iris spring 408 is attached to the second side of the irisplate 409 near its outside diameter. The iris spring 408 is alsoattached to the output stator 435 of the second variator 422. The irisspring 408 applies a resilient force that resists rotation of the irisplate 409 from tension applied by the iris cable 406. When tension fromthe iris cable 406 is released, the iris spring 408 returns the irisplate 409 to its at rest position. Depending upon the application of thetransmission 1700, the iris plate 409 can be configured so that when theiris cable 406 is pulled the iris plate 409 shifts the transmission 1700to a higher transmission ratio, and when tension on the iris cable 406is released the iris spring 408 shifts the transmission 1700 to a lowratio. Alternatively, the iris plate 409 can be configured so that whenthe iris cable 406 is pulled the iris plate 409 shifts the transmission1700 to a lower ratio, and when tension on the iris cable 406 isreleased the iris spring 408 shifts the transmission 1700 to a highratio.

Referring to FIGS. 16 and 17, embodiments of the transmission 1700having two variators 420, 422 require a high degree of accuracy in thealignment of the additional rolling elements of the transmission 1700.All of the rolling elements must be aligned with one another orefficiency will suffer and the lifespan of the transmission 1700 will bereduced. During assembly, the input disc 34, the output disc 430, thesecond input disc 431, and the idler assemblies 402 are aligned on thesame longitudinal axis. Additionally, the cage 410, which in theseembodiments consist of two cages 89 joined by the output stators 435 aspreviously described, must also be aligned on the longitudinal axis toaccurately position the ball/leg assemblies 403. To accomplish thissimply and accurately, all rolling elements are positioned relative tothe input shaft 425. A first input stator bearing 440 and a second inputstator bearing 444 are positioned in the bores of the input stators 440,444 and over the input shaft 425 to help align the cage 410. An outputstator bearing 442 positioned in the bore of the output stators 435 andover the input shaft 425 also aligns the cage 410. A first guide bearing441 is positioned in the bore of the first shift guide 13 b and over theinput shaft 425 and a second guide bearing 443 is positioned in the boreof the second shift guide 13 b and over the input shaft 425 to align thefirst and second idler assemblies 402.

Referring to FIGS. 18 and 19, the cage 410 is attached to the case 423with the previously described case connectors 383 that fit into caseslots 421. The case slots 421 are longitudinal grooves in the case 423that extend to the input side of the case 423, the side of the case 423that is open. In the illustrated embodiment, the case is mostly closedon the output side, which is not shown in FIG. 19, but is open on theinput side and has a mounting flange extending radially from the otherwise cylindrical body of the case 423 that case holes 424 for mountingthe case 423. During assembly, the transmission 1700 can be insertedinto the case 423 where the case connecters 383 are aligned in the caseslots 421 in order to resist torque applied to the cage 410 and preventthe cage 410 from rotating. Case connector holes 412 in the case 423allow fasteners to be inserted into corresponding holes in the caseconnectors 383 to fasten the cage 410 to the case 423.

FIG. 22 illustrates an alternate embodiment of the cage 470 of thetransmission 1700. To reduce manufacturing costs, it is sometimespreferable to minimize the number of different parts that aremanufactured and to design parts that can be inexpensively producedusing mass production techniques. The illustrated cage 470 uses fourdifferent parts of low cost design and common fasteners to assemble thevarious components. The stators 472 are generally flat disc shapedpieces with multiple radial grooves extending radially outward from neara central bore through which the input shaft 425 rotates. The ball axles(item 433 of FIG. 17) extend through the grooves on the stators 472.Multiple holes 471 surrounding the central bore of the stators 472provide for fastening the stators 472 to other components. There arefour stators 472, which in this embodiment are all similar to oneanother, forming part of the cage 470. Two input stators 472 are at eachend of the cage 470 and two output stators 472 are near the center ofthe cage 472, which are rigidly attached to each other with a statorbridge 477.

Still referring to the embodiment illustrated in FIG. 22, the statorbridge 477 is a disc shaped part with a central bore and through holespositioned between the inside diameter and the outside diameter of thestator bridge 477. The holes in the stator bridge 477 are complimentaryto the holes on the stators 472 to allow fastening of the stators 472 tothe stator bridge 477. The iris plate 409 (not shown) is locatedradially outside of the stator bridge 477 and axially between the outputstators 472. In some embodiments, the stator bridge 477 is slightlythicker than the iris plate 409 to allow freedom of rotation of the irisplate 409, while in yet other embodiments, bearings are located betweenthe output stators 472 and the iris plate 409, as well as between thestator bridge 477 and the iris plate 409. The outside diameter of thestator bridge 477, therefore serves to locate the inside diameter andset the axis of the iris plate 409.

Spacers 473 join the input stators 472 to the output stators 472. In oneembodiment, the spacers 473 are made from a flat material, such as sheetor plate metal, and are then formed to produce their unique shape, whichserves several purposes. The spacers 473, in general, are flatrectangular sheets with holes 475 formed in their centers and havingperpendicular extensions on each end. The spacers 473 set the correctdistance between the stators 472, form the structural frame of the cage470 to prevent the balls 1 from orbiting the longitudinal axis of thetransmission 1700, align the stator holes with respect to one another sothat the centers of the stators 472 are in alignment and the angularorientation of the stators 472 is the same, prevent the cage 470 fromtwisting or cocking, and provide rolling concave surfaces 479 on whichthe stator wheels 30 roll. Each spacer 473 is formed with its two endsbent out of plane with the rest of the spacer to form the mounting areas480 and curved surfaces 479 of the cage 470. The spacers 473 havemounting holes 481 on the sides where they contact the stators 472 whichline up with corresponding holes on the stators 472 to allow fasteningof the spacers 473 to the stators 472. The hole 475 near the center ofthe spacer 473 provides clearance for the ball 1.

In one embodiment, there are two spacers 473 for each ball 1 althoughmore or fewer spacers 473 can be used. Each spacer 473 is paired back toback with another in a mirror image to form an I-beam shape. In oneembodiment, rivets 476 may be used to connect the spacers 473 to thestators 472 and to connect the stators 472 to the stator bridge 477. Therivets 476 are tightly pressed into the holes of the stators 472, thespacers 473 and the stator bridge 477 during assembly. Only two rivets476 are illustrated in FIG. 22, but all can use the same design. Thespacers 473 used in the first variator 420 also have case connectors474, which generally extend radially outward from the spacers 473 andthen bend generally perpendicularly. The case connectors 474, of someembodiments are made from a flat material such as sheet metal, which isstamped and then formed into the final shape. The case connectors 474can be made integral with or rigidly attached to the spacers 473 andextend radially to the case 423 between the input disc 34 and the outputdisc 430. In some embodiments, the case connectors 474 are formed aspart of the spacers 473 during the manufacturing process of the spacers473. Case connector holes 478 in the perpendicular ends of the caseconnectors 474 line up with corresponding case connector holes (item 412of FIG. 19) so the cage 470 can be anchored to the case 423 withstandard fasteners.

The design illustrated in FIG. 22 incorporates stator discs 472 that aresubstantially flat and that can be manufactured utilizing asubstantially flat sheet of rigid material. Additionally, the spacers473 with and without the case connectors 474 are also substantially flatand can be formed from flat sheets of material, although in manyembodiments the perpendicular ends of the case connectors 474, themounting areas 480 and the curved surfaces 480 are formed in subsequentbending steps. The stator discs 472 and spacers can be produced from anyof a number of inexpensive manufacturing techniques such as stamping,fine blanking, or any other such technique known in the industry. Thestator discs 472 and spacers 473 of this design can be made from thin orsheet metal, plastic, ceramic, wood or paper products or any othermaterial. As described above with respect to FIG. 12, the illustrateddesign allows for significant reduction in the cost of materials andmanufacturing of these otherwise relatively expensive components to asuitably high tolerance. Additionally, although the embodimentillustrated in FIG. 22 represents a dual-cavity design for atransmission, the components manufactured through these inexpensivemanufacturing processes can be used for a single cavity design of thecage 470 as well. As an example, two illustrated stators discs 472 canbe attached to the spacers 473 having the case connectors 474 to theright of FIG. 22 to produce a single cavity design for use with theembodiments described herein.

FIG. 23 illustrates an embodiment of a ball 1 for use with thetransmissions 100, 1700 of FIG. 1 and FIG. 17. This ball 1 has helicalgrooves 450 that pump lubricant through the ball 1. In one embodiment,two helical grooves 450 are used that begin at one end of the hole inthe ball 1 and continue through to the other end of the hole. Thehelical grooves 450 transport lubricant through the ball 1 to removeheat and provide lubrication between the ball 1 and the ball axles 3,433 in order to improve efficiency and to improve the lifespan of thetransmission 100, 1700.

FIG. 24 illustrates an alternate leg 460 of the ball/leg assembly 403 ofFIG. 5. The leg 460 is simplified, as compared to the leg 2 illustratedin FIG. 5, and does not have stator wheels 30, a stator wheel pin 31, aguide wheel 21, or a guide wheel pin 22. The leg 460 has a convexsurface on a first leg side 463 that faces away from the ball 1, whichfits into a corresponding concave groove (not shown) on a respectivestator 80. On a second leg side 465 that faces the ball 1, the leg 460is concaved and has a convex curve near its radially inward end thatforms a leg cam 466, which contacts and is positioned axially andradially by the surfaces of the shift guides 13. Transverse andlongitudinal lubrication ports 462, 464, respectively, allow forlubrication to be fed into the leg and transported to different areas.Lubrication is used to cool the leg and other parts of the transmission100, 1700 and also to minimize friction where the leg contacts the shiftguide 13 and the stator 80. It should be noted that additional ports canbe drilled or formed in the leg 460 to direct lubrication to other areasand that any of the port openings may be used as an inlet for thelubrication. The longitudinal port 464 is an aperture running throughthe length of the leg 460, generally in the center and extending throughthe bottom and also through the ball axle bore 461 at the top of eachleg 460. The transverse port 462 is a blind hole formed approximatelyperpendicular to the longitudinal port 464 and extends out and beyondthe first leg side 463. In some embodiments, as illustrated, thetransverse port 462 intersects with longitudinal port 464 and terminatesand does not penetrate the second leg side 465. In some embodimentswhere the transverse port 462 intersects with the longitudinal port 464,lubricant can enter at the opening of the transverse port 462 and thenbe transported through port 464.

In some embodiments, the ball axles 3, 433 are press fit in the ball 1and rotate with the ball 1. The ball axles 3, 433 rotate inside the ballaxle bores 461 and in the rollers 4. Lubricant flows through the top ofthe leg 460 into the ball axle bore 461 where it provides a fluid layerto reduce friction.

Referring to FIGS. 25-27, a graphical method for approximating theconvex curve 97 on a shift guide 13 is disclosed. For the purpose ofsimplicity, the idler 18, the idler bearings 17, and the shift guides 13are combined to simplify the analysis and illustration of the correctconvex curves 97 of one embodiment of the shift guides 13. For thepurpose of this analysis and description, the following assumptions aremade:

-   -   1. The center of the ball 1 is fixed such that the ball 1 can        rotate about its axis and such that its axis can rotate, but the        ball 1 can have no displacement.    -   2. The ball 1, ball axle 3, 433, legs 2, and guide wheels 21        rotate as a rigid body.    -   3. The idler 18 can only move in the x direction.    -   4. The perimeter surface of the idler 18 is tangent to the        circumference of the ball 1.    -   5. The sides of the shift guides 13 are tangent to the        circumference of the guide wheels 21.    -   6. Angular rotation of the ball 1 causes linear movement of the        shift guide 13, and vice-versa.    -   7. When the ball axle 3, 433 is horizontal or parallel to the        longitudinal axis 11, the point of contact of each guide wheel        21 and its respective shift guide 13 is at the start of the        convex curve 97 where the vertical wall on the shift guide 13        transitions to the convex curve 97. When the ball 1 is tilted,        only one guide wheel 21 contacts the convex curve 97; the other        guide wheel 21 contacting the vertical wall of its shift guide        13.

The goal of this analysis is to find the approximate coordinates of thepoint where the guide wheel 21 contacts the convex curve 97 on the shiftguide 13 as a function of the angle of tilt of the axle of the ball 1.If these coordinates are plotted for various ball axle 3, 433 angles, acurve can be fit through the coordinate points that follow the path ofthe guide wheel 21/shift guide 13 contact points throughout the shiftingrange.

The coordinates begin at the original position of the guide wheel21/shift guide 13 contact (xo, yo) when the angle of rotation is zero,and then at each incremental angular change during the tilting of theball 1. By comparing these coordinates, the position of the guide wheel21/shift guide 13 contact (xn, yn) as a function of the angle of ball 1tilt (theta) can be determined.

From FIGS. 25 and 26, the known variables are:

-   -   1. H1: the vertical distance from the center of the ball 1 to        the center of the guide wheels 21.    -   2. H2: the sum of the ball 1 radius and the idler 18 radius.    -   3. W: the horizontal distance from the center of the ball 1 to        the center of the guide wheels 21.    -   4. rw: the guide wheel 21 radius.

From these known variables, the following relations can be identified:R 1=[(W−rw)² +H 1 ²]{circumflex over ( )}(½)  (1)Phi=TAN⁻¹](W−rw)/H 1]  (2)xo=w−rw  (3)yo=H 1−H 2  (4)BETA=TAN⁻¹(H 1/W)  (5)R 2=[H 1 ² +W ²]{circumflex over ( )}(½)  (6)

At this point, assume the ball 1 is tilted by angle, THETA, which causesthe shift guide 13 to move in the x direction (see FIG. 26). From this,the following can be found:Nu=90°−BETA−THETA  (7)x 2=R 2* SIN (Nu)  (8)x 3=x 2−rw  (9)x_shift guide=xo−x 3  (10)

This is the x distance the shift guide 13 moves for a given THETA.x 4=R 1* SIN (Phi+THETA)  (11)x_guide wheel=x 4−xo  (12)

This is the x distance the guide wheel 21 moves for a given THETA.

At this point, it is convenient to define an x′−y′ origin at the centerof the idler 18. This is useful for plotting the guide wheel 21/shiftguide 13 contact coordinates.x 1 =xo−(x_shift guide−x_guide wheel)  (13)

By combining Equations (10), (12), and (13),x 1=x 4+x 3 −xo  (14)

This is the x′ position of the guide wheel 21/shift guide 13 contact.

Finding the y′ position of the guide wheel 21/shift guide 13 contact isrelatively simple,y 2=R 1* COS (Phi+THETA)  (15)y 1=H 2 −y 2  (16)

This is the y′ position of the guide wheel 21/shift guide 13 contact.

Therefore, x1 and y1 can be determined and then plotted for variousvalues of THETA. This is shown graphically in FIG. 27. With thecoordinates in place, it is a simple matter for most CAD programs to fita curve through them. Methods of curve fitting can include any suitablealgorithm, such as for example linear regression, to determine theappropriate curve for such a relationship; although a direct functionderived from the relationships described above can be developed as well.

Referring now to FIGS. 1, 7, and 28, the transmission 100 can be used asa continuously variable planetary gearset 500. With reference to FIGS. 1and 7, in such embodiments where the cage 89 is free to rotate about thelongitudinal axis 11, the idler 18 functions as a sun gear, the balls 1act as planet gears, the cage 89 holds the balls 1 and functions as aplanet carrier, the input disc 34 is a first ring gear, and the outputdisc 101 is a second ring gear. Each ball 1 contacts the input disc 34,the output disc 101, and the idler 18 and is carried or held in radialposition by the cage 89.

FIG. 28 is a skeleton drawing, or a schematic view, of a planetarygearset 500 where, for simplicity, only the top half of the planetarygearset 500 is shown. The drawing is cut off at the centerline of theplanetary gearset 500, or on the longitudinal axis 11 of thetransmission 100. The line of contact formed around each of the balls 1by the output disc 101 forms a variable rolling diameter that allowsthat portion of each of the balls 1 to function as a first planet gear501. The contact between the balls 1 and idler 18 create a variablerolling diameter, which allows that portion of each of the balls 1 tofunction as a second planet gear 502. The contact between the balls 1and input disc 34 create a variable rolling diameter, which allows thatportion of the balls 1 to function as a third planet gear 503.

In embodiments of the planetary gear set 500, those of skill in the artwill recognize that various radial and thrust bearings canadvantageously be utilized to maintain the positions of the input disc34, output disc 101 and cage 89 with respect to one another. Those ofskill in the art will also recognize that solid or hollow shafts can beutilized and attached to the input disc 34, the output disc 101, thecage 89 and/or the idler 18 as appropriate to fulfill the functionsdescribed herein and such modifications are well within the skill ofthose in the field of rotational power transmission.

Referring now to FIGS. 29-31, the respective diameters of the firstplanet gear 501, second planet gear 502, and third planet gear 503 canbe changed by shifting the transmission 100. FIG. 29 shows thetransmission 100 with the first and third planet gears 501, 503 of equaldiameter, and the second planet gear 502 at its maximum diameter. Bytilting the balls 1 as previously described, the diameters of the planetgears 501, 502, 503 change, varying the input to output speed of thetransmission 1700. FIG. 30 shows the balls 1 tilted so that the firstplanet gear 501 is increased in diameter, and the second and thirdplanet gears 502 and 503 are decreased in diameter. FIG. 31 shows theballs tilted so that the third planet gear 503 is increased in diameterand the first and second planet gears 501 and 502 are decreased indiameter.

There are many different speed combinations possible by altering thesource of torque between the input disc 34, the idler 18, and/or thecage 89. Additionally, some embodiments utilize more than one input. Forexample, the input disc 34 and the cage 89 can both provide input torqueand can rotate at the same speed or different speeds. One or moresources of input torque can be capable of variable speed to increase theratio possibilities of the transmission 100. A list is provided below ofsome of the combinations available by using the transmission 100 as aplanetary gearset. In this list, a source of input torque, or an“input,” is coded with an “I”, an output is coded with an “O”, acomponent that is fixed such that it does not rotate about thelongitudinal axis 11 is coded with an “F”, and if a component is allowedto rotate freely, it is coded with an “R.” “Single In/Single Out” isused to indicate that there is one input and one output, “Dual In/SingleOut” is used to indicate that there are two inputs and one output,“Single ln/Dual Out” is used to indicate that there is one input and twooutputs, “Dual In/Dual Out” is used to indicate that there are twoinputs and two outputs, “Triple In/Single Out” is used to indicate thatthere are three inputs and one output, and “Single In/Triple Out” isused to indicate that there is one input and three outputs. Input DiscCage Output Configuration (34) Idler (18) (89) Disc (101) SingleIn/Single Out F I F O Single In/Single Out R I F O Single In/Single OutF I R O Single In/Single Out R I R O Single In/Single Out F I O F SingleIn/Single Out R I O F Single In/Single Out F I O R Single In/Single OutR I O R Single In/Single Out I R F O Single In/Single Out I F R O SingleIn/Single Out I F F O Single In/Single Out I R R O Single In/Single OutI F O F Single In/Single Out I F O R Single In/Single Out I R O F SingleIn/Single Out I R O R Single In/Single Out F F I O Single In/Single OutF R I O Single In/Single Out R F I O Single In/Single Out R R I O SingleIn/Single Out F O I F Single In/Single Out R O I F Single In/Single OutF O I R Single In/Single Out R O I R Dual In/Single Out I I F O DualIn/Single Out I I R O Dual In/Single Out I I O F Dual In/Single Out I IO R Dual In/Single Out I O I F Dual In/Single Out I O I R Dual In/SingleOut I F I O Dual In/Single Out I R I O Dual In/Single Out F I I O DualIn/Single Out R I I O Single In/Dual Out I O F O Single In/Dual Out I OR O Single In/Dual Out I F O O Single In/Dual Out I R O O Single In/DualOut I O O F Single In/Dual Out I O O R Single In/Dual Out F I O O SingleIn/Dual Out R I O O Single In/Dual Out F O I O Single In/Dual Out R O IO Dual In/Dual Out I I O O Dual In/Dual Out I O I O Triple In/Single OutI I I O Triple In/Single Out I I O I Single In/Triple Out I O O O

Referring to FIG. 32, the transmission 100 may also be combined througha parallel power path with a planetary gearset 505 to produce many morespeed combinations. A typical planetary gearset 505 is comprised of asun gear in the center, multiple planet gears distributed around andengaging the sun gear that are all rotatably attached at theirrespective centers to a planet carrier, often simply referred to as thecarrier, and a ring gear surrounding and engaging the planet gears. Byswitching the source of input torque and the output among the sun gear,carrier, and ring gear, many speed combinations can be obtained. Theplanetary gearset 505 combined with the transmission 100 provides for avery high number of speed combinations and in some cases an infinitelyvariable transmission can be obtained. In FIG. 32, the torque input ofthe transmission 100 is coupled both to the input disc 34 and to a firstgear 506, which is generally coaxial with input disc 34 and contacts androtates a second gear 509 to drive the parallel power path. The basicconfiguration of coupling both the input disc 34 of the transmission100, or CVT 100, and the input of a parallel power path to a prime moveror other torque source such as a motor or other powering device, istermed “Input Coupled.” By varying the diameters of the first gear 506and the second gear 509, the input speed to the parallel power path canbe varied. The second gear 509 is attached to and rotates a gear shaft508, which in some embodiments rotates a gearbox 507. The gearbox 507,implemented as a design option in such embodiments, can further vary therotation speed of the parallel power path and can be a conventionalgeared transmission. The gearbox 507 rotates a gearbox shaft 511, whichrotates a third gear 510. In embodiments not utilizing the gearbox 507,the gear shaft 508 drives the third gear 510. The third gear 510 drivesthe sun, carrier, or ring of the planetary gearset 505 and is of adiameter designed to create a desirable speed/torque ratio.Alternatively, the third gear 510 can be eliminated and the gearboxshaft 508 can rotate the sun, carrier, or ring of the planetary gearset505 directly. The planetary gearset 505 also has an input from the CVT100 output, which drives another of the sun, carrier or ring.

In the following table, titled “Input Coupled,” many, if not all, of thevarious input and output combinations that are possible with the basicarrangement as just described above are identified. In this table, “IT”represents the source of input torque into the CVT 100, “O” representsthe component of the CVT coupled to the planetary gearset 505, “I1”represents the planetary gearset 505 component coupled to the CVT 100output, “OV” represents the component of the planetary gearset 505 thatis connected to the output of the vehicle or machine, “F” represents acomponent of the planetary gearset 505 or the transmission 100 that isfixed so as not to rotate about its axis, “I2” represents a componentcoupled to the parallel path, which is the third gear 509, and “R”represents a component that is free to rotate about its axis andtherefore does not drive another component. For this table and the tablethat follows, entitled “Output Coupled,” it is assumed that the ringgear is the only planeteary gearset 505 component that is being fixed,in order to reduce the overall number of tables that have to be providedherein. The sun gear or the planet carrier can also be fixed withcorresponding input and output combinations for the other components andthose combinations are not provided herein in order to reduce the sizeof this description, but are easily determined based upon the followingtwo tables. Input Coupled CVT Planetary Gearset IT = Input I1 = Coupledto CVT Output O = Output to planetary input OV = Output to vehicle/loadF = Fixed to ground F = Fixed to ground R = Rolling (free) I2 = Coupledto parallel path Input Output Disc Idler Cage Disc Variator (34) (18)(89) (101) Ring Carrier Sun Single In/ F IT F O I1 I2 OV Single Out F ITF O OV I1 I2 F IT F O I2 OV I1 F IT F O F I1 I2, OV F IT F O F I2, OV I1F IT F O F I2 I1, OV F IT F O F I1, OV I2 Single In/ R IT F O I1 I2 OVSingle Out R IT F O OV I1 I2 R IT F O I2 OV I1 R IT F O F I1 I2, OV R ITF O F I2, OV I1 R IT F O F I2 I1, OV R IT F O F I1, OV I2 Single In/ FIT R O I1 I2 OV Single Out F IT R O OV I1 I2 F IT R O I2 OV I1 F IT R OF I1 I2, OV F IT R O F I2, OV I1 F IT R O F I2 I1, OV F IT R O F I1, OVI2 Single In/ R IT R O I1 I2 OV Single Out R IT R O OV I1 I2 R IT R O I2OV I1 R IT R O F I1 I2, OV R IT R O F I2, OV I1 R IT R O F I2 I1, OV RIT R O F I1, OV I2 Single In/ F IT O F I1 I2 OV Single Out F IT O F OVI1 I2 F IT O F I2 OV I1 F IT O F F I1 I2, OV F IT O F F I2, OV I1 F IT OF F I2 I1, OV F IT O F F I1, OV I2 Single In/ R IT O F I1 I2 OV SingleOut R IT O F OV I1 I2 R IT O F I2 OV I1 R IT O F F I1 I2, OV R IT O F FI2, OV I1 R IT O F F I2 I1, OV R IT O F F I1, OV I2 Single In/ F IT O RI1 I2 OV Single Out F IT O R OV I1 I2 F IT O R I2 OV I1 F IT O R F I1I2, OV F IT O R F I2, OV I1 F IT O R F I2 I1, OV F IT O R F I1, OV I2Single In/ R IT O R I1 I2 OV Single Out R IT O R OV I1 I2 R IT O R I2 OVI1 R IT O R F I1 I2, OV R IT O R F I2, OV I1 R IT O R F I2 I1, OV R IT OR F I1, OV I2 Single In/ IT R F O I1 I2 OV Single Out IT R F O OV I1 I2IT R F O I2 OV I1 IT R F O F I1 I2, OV IT R F O F I2, OV I1 IT R F O FI2 I1, OV IT R F O F I1, OV I2 Single In/ IT F R O I1 I2 OV Single OutIT F R O OV I1 I2 IT F R O I2 OV I1 IT F R O F I1 I2, OV IT F R O F I2,OV I1 IT F R O F I2 I1, OV IT F R O F I1, OV I2 Single In/ IT F F O I1I2 OV Single Out IT F F O OV I1 I2 IT F F O I2 OV I1 IT F F O F I1 I2,OV IT F F O F I2, OV I1 IT F F O F I2 I1, OV IT F F O F I1, OV I2 SingleIn/ IT R R O I1 I2 OV Single Out IT R R O OV I1 I2 IT R R O I2 OV I1 ITR R O F I1 I2, OV IT R R O F I2, OV I1 IT R R O F I2 I1, OV IT R R O FI1, OV I2 Single In/ IT F O F I1 I2 OV Single Out IT F O F OV I1 I2 IT FO F I2 OV I1 IT F O F F I1 I2, OV IT F O F F I2, OV I1 IT F O F F I2 I1,OV IT F O F F I1, OV I2 Single In/ IT F O R I1 I2 OV Single Out IT F O ROV I1 I2 IT F O R I2 OV I1 IT F O R F I1 I2, OV IT F O R F I2, OV I1 ITF O R F I2 I1, OV IT F O R F I1, OV I2 Single In/ IT R O F I1 I2 OVSingle Out IT R O F OV I1 I2 IT R O F I2 OV I1 IT R O F F I1 I2, OV IT RO F F I2, OV I1 IT R O F F I2 I1, OV IT R O F F I1, OV I2 Single In/ ITR O R I1 I2 OV Single Out IT R O R OV I1 I2 IT R O R I2 OV I1 IT R O R FI1 I2, OV IT R O R F I2, OV I1 IT R O R F I2 I1, OV IT R O R F I1, OV I2Single In/ F F IT O I1 I2 OV Single Out F F IT O OV I1 I2 F F IT O I2 OVI1 F F IT O F I1 I2, OV F F IT O F I2, OV I1 F F IT O F I2 I1, OV F F ITO F I1, OV I2 Single In/ F R IT O I1 I2 OV Single Out F R IT O OV I1 I2F R IT O I2 OV I1 F R IT O F I1 I2, OV F R IT O F I2, OV I1 F R IT O FI2 I1, OV F R IT O F I1, OV I2 Single In/ R F IT O I1 I2 OV Single Out RF IT O OV I1 I2 R F IT O I2 OV I1 R F IT O F I1 I2, OV R F IT O F I2, OVI1 R F IT O F I2 I1, OV R F IT O F I1, OV I2 Single In/ R R IT O I1 I2OV Single Out R R IT O OV I1 I2 R R IT O I2 OV I1 R R IT O F I1 I2, OV RR IT O F I2, OV I1 R R IT O F I2 I1, OV R R IT O F I1, OV I2 Single In/F O IT F I1 I2 OV Single Out F O IT F OV I1 I2 F O IT F I2 OV I1 F O ITF F I1 I2, OV F O IT F F I2, OV I1 F O IT F F I2 I1, OV F O IT F F I1,OV I2 Single In/ R O IT F I1 I2 OV Single Out R O IT F OV I1 I2 R O IT FI2 OV I1 R O IT F F I1 I2, OV R O IT F F I2, OV I1 R O IT F F I2 I1, OVR O IT F F I1, OV I2 Single In/ F O IT R I1 I2 OV Single Out F O IT R OVI1 I2 F O IT R I2 OV I1 F O IT R F I1 I2, OV F O IT R F I2, OV I1 F O ITR F I2 I1, OV F O IT R F I1, OV I2 Single In/ R O IT R I1 I2 OV SingleOut R O IT R OV I1 I2 R O IT R I2 OV I1 R O IT R F I1 I2, OV R O IT R FI2, OV I1 R O IT R F I2 I1, OV R O IT R F I1, OV I2

Referring to the embodiment illustrated in FIG. 33, the source of torqueinput drives the planetary gearset 505, which is coupled as an input tothe CVT 100. One or more components of the CVT 100 are coupled to aparallel power path and to the output of the transmission. The parallelpower path in this embodiment is as follows: a component of theplanetary gearset 505, either the sun, the carrier, or the ring, mesheswith a third gear 510, which rotates the gear shaft 508, which in turndrives the previously described gearbox 507. The gearbox 507 rotates thegearbox shaft 511, which rotates the second gear 509, which in turndrives the first gear 506. The first gear 506 is then mounted on theoutput shaft of the transmission, which is also coupled to the output ofthe CVT 100. In this embodiment, the planetary gearset 505 is coupled tothe source of torque to the transmission and then provides torque toboth the parallel path and the CVT 100 and the torque from both of thesepaths is coupled at the output of the vehicle or equipment. If theplanetary gearset 505 is coupled thusly to provide torque to the CVT 100and to the fixed ratio parallel path, and both paths are coupled at theoutput, such as in a drive shaft, wheel, or other loaded device, theconfiguration can be referred to as “Output Coupled.” In this basicconfiguration, the planetary gearset 505 combined with the CVT 100provides for a very high number of speed combinations and in some casesan infinitely variable transmission can be obtained.

In the following table, titled “Output Coupled,” many if not all of thepossible combinations of the basic arrangement shown in FIG. 33 areprovided and described. In this table, for the planetary gearset 505,“O1” refers to the component of the planetary gearset 505 coupled to theCVT 100, “I” refers to the input from the engine, human, or whateversource, “F” refers to a component that is fixed so as not to rotateabout its own axis, and “O2” refers to the component coupled to theparallel path, via planetary gear 510. For the CVT 100, “I” refers tothe component that is coupled to the planetary gearset 505, “O” refersto the component that is coupled to the output of the vehicle ormachine, “F” refers to a fixed component as just described, and “R”refers to a component that is free to rotate about its axis, andtherefore does not drive any other component. Output Coupled PlanetaryGearset CVT O1 = Coupled to CVT Input I = Input from Planetary output I= Input from engine O = Output to vehicle/load F = Fixed to ground F =Fixed to ground O2 = Coupled to parallel path R = Rolling (free) InputOutput Disc Idler Cage Disc Variator Ring Carrier Sun (34) (18) (89)(101) Single In/ I O1 O2 F I F O Single Out O2 I O1 F I F O O1 O2 I F IF O F I, O1 O2 F I F O F O2 I, O1 F I F O F I, O2 O1 F I F O F O1 I, O2F I F O Single In/ I O1 O2 R I F O Single Out O2 I O1 R I F O O1 O2 I RI F O F I, O1 O2 R I F O F O2 I, O1 R I F O F I, O2 O1 R I F O F O1 I,O2 R I F O Single In/ I O1 O2 F I R O Single Out O2 I O1 F I R O O1 O2 IF I R O F I, O1 O2 F I R O F O2 I, O1 F I R O F I, O2 O1 F I R O F O1 I,O2 F I R O Single In/ I O1 O2 R I R O Single Out O2 I O1 R I R O O1 O2 IR I R O F I, O1 O2 R I R O F O2 I, O1 R I R O F I, O2 O1 R I R O F O1 I,O2 R I R O Single In/ I O1 O2 F I O F Single Out O2 I O1 F I O F O1 O2 IF I O F F I, O1 O2 F I O F F O2 I, O1 F I O F F I, O2 O1 F I O F F O1 I,O2 F I O F Single In/ I O1 O2 R I O F Single Out O2 I O1 R I O F O1 O2 IR I O F F I, O1 O2 R I O F F O2 I, O1 R I O F F I, O2 O1 R I O F F O1 I,O2 R I O F Single In/ I O1 O2 F I O R Single Out O2 I O1 F I O R O1 O2 IF I O R F I, O1 O2 F I O R F O2 I, O1 F I O R F I, O2 O1 F I O R F O1 I,O2 F I O R Single In/ I O1 O2 R I O R Single Out O2 I O1 R I O R O1 O2 IR I O R F I, O1 O2 R I O R F O2 I, O1 R I O R F I, O2 O1 R I O R F O1 I,O2 R I O R Single In/ I O1 O2 I R F O Single Out O2 I O1 I R F O O1 O2 II R F O F I, O1 O2 I R F O F O2 I, O1 I R F O F I, O2 O1 I R F O F O1 I,O2 I R F O Single In/ I O1 O2 I F R O Single Out O2 I O1 I F R O O1 O2 II F R O F I, O1 O2 I F R O F O2 I, O1 I F R O F I, O2 O1 I F R O F O1 I,O2 I F R O Single In/ I O1 O2 I F F O Single Out O2 I O1 I F F O O1 O2 II F F O F I, O1 O2 I F F O F O2 I, O1 I F F O F I, O2 O1 I F F O F O1 I,O2 I F F O Single In/ I O1 O2 I R R O Single Out O2 I O1 I R R O O1 O2 II R R O F I, O1 O2 I R R O F O2 I, O1 I R R O F I, O2 O1 I R R O F O1 I,O2 I R R O Single In/ I O1 O2 I F O F Single Out O2 I O1 I F O F O1 O2 II F O F F I, O1 O2 I F O F F O2 I, O1 I F O F F I, O2 O1 I F O F F O1 I,O2 I F O F Single In/ I O1 O2 I F O R Single Out O2 I O1 I F O R O1 O2 II F O R F I, O1 O2 I F O R F O2 I, O1 I F O R F I, O2 O1 I F O R F O1 I,O2 I F O R Single In/ I O1 O2 I R O F Single Out O2 I O1 I R O F O1 O2 II R O F F I, O1 O2 I R O F F O2 I, O1 I R O F F I, O2 O1 I R O F F O1 I,O2 I R O F Single In/ I O1 O2 I R O R Single Out O2 I O1 I R O R O1 O2 II R O R F I, O1 O2 I R O R F O2 I, O1 I R O R F I, O2 O1 I R O R F O1 I,O2 I R O R Single In/ I O1 O2 F F I O Single Out O2 I O1 F F I O O1 O2 IF F I O F I, O1 O2 F F I O F O2 I, O1 F F I O F I, O2 O1 F F I O F O1 I,O2 F F I O Single In/ I O1 O2 F R I O Single Out O2 I O1 F R I O O1 O2 IF R I O F I, O1 O2 F R I O F O2 I, O1 F R I O F I, O2 O1 F R I O F O1 I,O2 F R I O Single In/ I O1 O2 R F I O Single Out O2 I O1 R F I O O1 O2 IR F I O F I, O1 O2 R F I O F O2 I, O1 R F I O F I, O2 O1 R F I O F O1 I,O2 R F I O Single In/ I O1 O2 R R I O Single Out O2 I O1 R R I O O1 O2 IR R I O F I, O1 O2 R R I O F O2 I, O1 R R I O F I, O2 O1 R R I O F O1 I,O2 R R I O Single In/ I O1 O2 F O I F Single Out O2 I O1 F O I F O1 O2 IF O I F F I, O1 O2 F O I F F O2 I, O1 F O I F F I, O2 O1 F O I F F O1 I,O2 F O I F Single In/ I O1 O2 R O I F Single Out O2 I O1 R O I F O1 O2 IR O I F F I, O1 O2 R O I F F O2 I, O1 R O I F F I, O2 O1 R O I F F O1 I,O2 R O I F Single In/ I O1 O2 F O I R Single Out O2 I O1 F O I R O1 O2 IF O I R F I, O1 O2 F O I R F O2 I, O1 F O I R F I, O2 O1 F O I R F O1 I,O2 F O I R Single In/ I O1 O2 R O I R Single Out O2 I O1 R O I R O1 O2 IR O I R F I, O1 O2 R O I R F O2 I, O1 R O I R F I, O2 O1 R O I R F O1 I,O2 R O I R

Referring to the embodiment illustrated in FIG. 32, the following table,titled “Input Coupled Dual Input Power paths,” shows combinations in abasic input coupled arrangement with two sources of torque input intothe planetary gearset 505. The reference letters provided in this tablerepresent the same components as they did in the previous table exceptthat for the planetary gearset 505, “I1” refers to the output of the CVT100 and “I2” is the component that is coupled to the parallel path,which in this case is the planetary gear 510. Input Coupled Dual InputPower paths CVT Planetary Gearset I = Input I1 = Coupled to CVT Output O= Output O = Output to vehicle/load F = Fixed to ground F = Fixed toground R = Rolling (free) I2 = Coupled to parallel path Input OutputDisc Idler Cage Disc Variator (34) (18) (89) (101) Ring Carrier Sun DualIn/ I I F O I1 I2 O Single Out I I F O O I1 I2 I I F O I2 O I1 I I F O FI1 I2, O I I F O F I2, O I1 I I F O F I2 I1, O I I F O F I1, O I2 DualIn/ I I R O I1 I2 O Single Out I I R O O I1 I2 I I R O I2 O I1 I I R O FI1 I2, O I I R O F I2, O I1 I I R O F I2 I1, O I I R O F I1, O I2 DualIn/ I I O F I1 I2 O Single Out I I O F O I1 I2 I I O F I2 O I1 I I O F FI1 I2, O I I O F F I2, O I1 I I O F F I2 I1, O I I O F F I1, O I2 DualIn/ I I O R I1 I2 O Single Out I I O R O I1 I2 I I O R I2 O I1 I I O R FI1 I2, O I I O R F I2, O I1 I I O R F I2 I1, O I I O R F I1, O I2 DualIn/ I O I F I1 I2 O Single Out I O I F O I1 I2 I O I F I2 O I1 I O I F FI1 I2, O I O I F F I2, O I1 I O I F F I2 I1, O I O I F F I1, O I2 DualIn/ I O I R I1 I2 O Single Out I O I R O I1 I2 I O I R I2 O I1 I O I R FI1 I2, O I O I R F I2, O I1 I O I R F I2 I1, O I O I R F I1, O I2 DualIn/ I F I O I1 I2 O Single Out I F I O O I1 I2 I F I O I2 O I1 I F I O FI1 I2, O I F I O F I2, O I1 I F I O F I2 I1, O I F I O F I1, O I2 DualIn/ I R I O I1 I2 O Single Out I R I O O I1 I2 I R I O I2 O I1 I R I O FI1 I2, O I R I O F I2, O I1 I R I O F I2 I1, O I R I O F I1, O I2 DualIn/ F I I O I1 I2 O Single Out F I I O O I1 I2 F I I O I2 O I1 F I I O FI1 I2, O F I I O F I2, O I1 F I I O F I2 I1, O F I I O F I1, O I2 DualIn/ R I I O I1 I2 O Single Out R I I O O I1 I2 R I I O I2 O I1 R I I O FI1 I2, O R I I O F I2, O I1 R I I O F I2 I1, O R I I O F I1, O I2

Still referring to the embodiment illustrated in FIG. 32, the followingtable, titled “Input Coupled Triple Input” refers to embodimentsutilizing three sources of input torque into the CVT 100. For thistable, the CVT 100 reference letters refer to the same components as inthe previous table and the planetary gearset 505 reference lettersrepresent the same components except for “I2,” which represents thecomponent that is coupled to the parallel path. Input Coupled TripleInput CVT Planetary Gearset I = Input I1 = Coupled to CVT Output O =Output O = Output to vehicle/load F = Fixed to ground F = Fixed toground R = Rolling (free) I2 = Coupled to parallel path Input OutputDisc Idler Cage Disc Variator (34) (18) (89) (101) Ring Carrier SunTriple In/ I I I O I1 I2 O Single Out I I I O O I1 I2 I I I O I2 O I1 II I O F I1 I2, O I I I O F I2, O I1 I I I O F I2 I1, O I I I O F I1, OI2 Triple In/ I I O I I1 I2 O Single Out I I O I O I1 I2 I I O I I2 O I1I I O I F I1 I2, O I I O I F I2, O I1 I I O I F I2 I1, O I I O I F I1, OI2

Referring now to the embodiment illustrated in FIG. 34, the parallelpath can be eliminated due to the unique arrangement of the embodimentsdescribed herein. The parallel path is now combined into a collineararrangement where various components of the CVT and the planetarygearset 505 are coupled to produce all of the combinations describedabove and below. In some embodiments, the planetary gearset 505 iscoupled to the input of the CVT 100 or, as illustrated in FIG. 34, itcan be coupled to the output of the CVT 100. The following table, titled“Input Coupled Dual Output Power paths” lists various combinations thatare available where there are two outputs from the CVT 100 into theplanetary gearset 505. The reference letters for the CVT 100 are thesame as the previous table and the planetary gearset 505 referenceletters represent the same components except for “I2,” which is nolonger coupled to the parallel path but is coupled to the second CVT 100output. Input Coupled Dual Output Power paths CVT Planetary Gearset I =Input I1 = Coupled to CVT Output O = Output O = Output to vehicle/load F= Fixed to ground R = Free to Roll R = Rolling (free) I2 = Coupled toCVT output Input Output Disc Idler Cage Disc Variator (34) (18) (89)(101) Ring Carrier Sun Single In/ I O F O I1 I2 O Dual Out I O F O O I1I2 I O F O I2 O I1 I O F O F I1 I2, O I O F O F I2, O I1 I O F O F I2I1, O I O F O F I1, O I2 Single In/ I O R O I1 I2 O Dual Out I O R O OI1 I2 I O R O I2 O I1 I O R O F I1 I2, O I O R O F I2, O I1 I O R O F I2I1, O I O R O F I1, O I2 Single In/ I R O O I1 I2 O Dual Out I R O O OI1 I2 I R O O I2 O I1 I R O O F I1 I2, O I R O O F I2, O I1 I R O O F I2I1, O I R O O F I1, O I2 Single In/ I F O O I1 I2 O Dual Out I F O O OI1 I2 I F O O I2 O I1 I F O O F I1 I2, O I F O O F I2, O I1 I F O O F I2I1, O I F O O F I1, O I2 Single In/ I O O F I1 I2 O Dual Out I O O F OI1 I2 I O O F I2 O I1 I O O F F I1 I2, O I O O F F I2, O I1 I O O F F I2I1, O I O O F F I1, O I2 Single In/ I O O R I1 I2 O Dual Out I O O R OI1 I2 I O O R I2 O I1 I O O R F I1 I2, O I O O R F I2, O I1 I O O R F I2I1, O I O O R F I1, O I2 Single In/ F I O O I1 I2 O Dual Out F I O O OI1 I2 F I O O I2 O I1 F I O O F I1 I2, O F I O O F I2, O I1 F I O O F I2I1, O F I O O F I1, O I2 Single In/ R I O O I1 I2 O Dual Out R I O O OI1 I2 R I O O I2 O I1 R I O O F I1 I2, O R I O O F I2, O I1 R I O O F I2I1, O R I O O F I1, O I2 Single In/ F O I O I1 I2 O Dual Out F O I O OI1 I2 F O I O I2 O I1 F O I O F I1 I2, O F O I O F I2, O I1 F O I O F I2I1, O F O I O F I1, O I2 Single In/ R O I O I1 I2 O Dual Out R O I O OI1 I2 R O I O I2 O I1 R O I O F I1 I2, O R O I O F I2, O I1 R O I O F I2I1, O R O I O F I1, O I2

For the preceding two tables, the transmissions described could beinverted to provide an inverted result for each combination, but suchreverse combinations are easily recognized and are not separatelydescribed herein for space considerations. For instance, for OutputCoupled Dual Output, the inverse of Input Coupled/Dual Input, note thateither planetary gearset 505 input could be coupled to either CVT 100output.

Still referring to the embodiment illustrated in FIG. 34, the followingtable titled “Input Coupled Dual-Dual” provides various combinationsavailable where there are two sources of torque input into the CVT 100and two outputs from the CVT 100 into the planetary gearset 505. InputCoupled Dual-Dual CVT Planetary Gearset I = Input I1 = Coupled to CVTOutput O = Output O = Output to vehicle/load F = Fixed to ground R =Free to Roll R = Rolling (free) I2 = Coupled to CVT output Input OutputDisc Idler Cage Disc Variator (34) (18) (89) (101) Ring Carrier Sun DualIn/ I I O O I1 I2 O Dual Out I I O O O I1 I2 I I O O I2 O I1 I I O O FI1 I2, O I I O O F I2, O I1 I I O O F I2 I1, O I I O O F I1, O I2 DualIn/ I O I O I1 I2 O Dual Out I O I O O I1 I2 I O I O I2 O I1 I O I O FI1 I2, O I O I O F I2, O I1 I O I O F I2 I1, O I O I O F I1, O I2

Still referring to FIG. 34, the following table, titled “InternallyCoupled Planetary on Output,” provides many if not all of thecombinations available when the planetary gearset 505 is coupleddirectly to components of the CVT 100. For the CVT 100, the referenceletters “O1” indicate a component that is coupled to “I1” of theplanetary gearset 505, “R” represents a component that is rolling freeor a second input, “F” represents a component that is rigidly attachedto a stationary component, such as a fixed casing or to a supportstructure for the transmission, and “O2” is coupled to “I2” of theplanetary gearset 505. For the planetary gearset 505, “I1” refers to acomponent that is coupled to a first output component of the CVT 100,“O” refers to a component providing the output to a vehicle or otherloaded device, “F” is fixed, and “I2” is coupled to a second CVT 100output component. It should be noted that for the combinationsillustrated in the following table, the input element could also becoupled to any one of the planetary elements with corresponding changesto the coupling arrangement of the other elements Internally CoupledPlanetary on Output CVT Planetary Gearset O1 = Coupled to Planetary I1I1 = Coupled to CVT Output R = Rolling free or 2nd Input O = Output tovehicle/load F = Fixed to ground F = Fixed to ground O2 = Coupled toPlanetary I2 I2 = Coupled to Second CVT Output Input Output Disc IdlerCage Disc Variator (34) (18) (89) (10) Ring Carrier Sun Single In/ I O1F O2 I1 I2 O Dual Out I O2 F O1 I2 I1 O I O1 F O2 O I1 I2 I O2 F O1 O I2I1 I O1 F O2 I2 O I1 I O2 F O1 I1 O I2 I O1 F O2 F I1 I2, O I O2 F O1 FI2, O I1 I O1 F O2 F I2, O I1 I O2 F O1 F I1 I2, O I O1 F O2 F I2 I1, OI O2 F O1 F I1, O I2 I O1 F O2 F I1, O I2 I O2 F O1 F I2 I1, O SingleIn/ I O1 R O2 I1 I2 O Dual Out I O2 R O1 I2 I1 O I O1 R O2 O I1 I2 I O2R O1 O I2 I1 I O1 R O2 I2 O I1 I O2 R O1 I1 O I2 I O1 R O2 F I1 I2, O IO2 R O1 F I2, O I1 I O1 R O2 F I2, O I1 I O2 R O1 F I1 I2, O I O1 R O2 FI2 I1, O I O2 R O1 F I1, O I2 I O1 R O2 F I1, O I2 I O2 R O1 F I2 I1, OSingle In/ I O1 O2 F I1 I2 O Dual Out I O2 O1 F I2 I1 O I O1 O2 F O I1I2 I O2 O1 F O I2 I1 I O1 O2 F I2 O I1 I O2 O1 F I1 O I2 I O1 O2 F F I1I2, O I O2 O1 F F I2, O I1 I O1 O2 F F I2, O I1 I O2 O1 F F I1 I2, O IO1 O2 F F I2 I1, O I O2 O1 F F I1, O I2 I O1 O2 F F I1, O I2 I O2 O1 F FI2 I1, O Single In/ I O1 O2 R I1 I2 O Dual Out I O2 O1 R I2 I1 O I O1 O2R O I1 I2 I O2 O1 R O I2 I1 I O1 O2 R I2 O I1 I O2 O1 R I1 O I2 I O1 O2R F I1 I2, O I O2 O1 R F I2, O I1 I O1 O2 R F I2, O I1 I O2 O1 R F I1I2, O I O1 O2 R F I2 I1, O I O2 O1 R F I1, O I2 I O1 O2 R F I1, O I2 IO2 O1 R F I2 I1, O Single In/ O1 O2 I F I1 I2 O Dual Out O2 O1 I F I2 I1O O1 O2 I F O I1 I2 O2 O1 I F O I2 I1 O1 O2 I F I2 O I1 O2 O1 I F I1 OI2 O1 O2 I F F I1 I2, O O2 O1 I F F I2, O I1 O1 O2 I F F I2, O I1 O2 O1I F F I1 I2, O O1 O2 I F F I2 I1, O O2 O1 I F F I1, O I2 O1 O2 I F F I1,O I2 O2 O1 I F F I2 I1, O Single In/ O1 O2 I R I1 I2 O Dual Out O2 O1 IR I2 I1 O O1 O2 I R O I1 I2 O2 O1 I R O I2 I1 O1 O2 I R I2 O I1 O2 O1 IR I1 O I2 O1 O2 I R F I1 I2, O O2 O1 I R F I2, O I1 O1 O2 I R F I2, O I1O2 O1 I R F I1 I2, O O1 O2 I R F I2 I1, O O2 O1 I R F I1, O I2 O1 O2 I RF I1, O I2 O2 O1 I R F I2 I1, O Single In/ I F O1 O2 I1 I2 O Dual Out IF O2 O1 I2 I1 O I F O1 O2 O I1 I2 I F O2 O1 O I2 I1 I F O1 O2 I2 O I1 IF O2 O1 I1 O I2 I F O1 O2 F I1 I2, O I F O2 O1 F I2, O I1 I F O1 O2 FI2, O I1 I F O2 O1 F I1 I2, O I F O1 O2 F I2 I1, O I F O2 O1 F I1, O I2I F O1 O2 F I1, O I2 I F O2 O1 F I2 I1, O Single In/ I R O1 O2 I1 I2 ODual Out I R O2 O1 I2 I1 O I R O1 O2 O I1 I2 I R O2 O1 O I2 I1 I R O1 O2I2 O I1 I R O2 O1 I1 O I2 I R O1 O2 F I1 I2, O I R O2 O1 F I2, O I1 I RO1 O2 F I2, O I1 I R O2 O1 F I1 I2, O I R O1 O2 F I2 I1, O I R O2 O1 FI1, O I2 I R O1 O2 F I1, O I2 I R O2 O1 F I2 I1, O Single In/ F O1 I O2I1 I2 O Dual Out F O2 I O1 I2 I1 O F O1 I O2 O I1 I2 F O2 I O1 O I2 I1 FO1 I O2 I2 O I1 F O2 I O1 I1 O I2 F O1 I O2 F I1 I2, O F O2 I O1 F I2, OI1 F O1 I O2 F I2, O I1 F O2 I O1 F I1 I2, O F O1 I O2 F I2 I1, O F O2 IO1 F I1, O I2 F O1 I O2 F I1, O I2 F O2 I O1 F I2 I1, O Single In/ R O1I O2 I1 I2 O Dual Out R O2 I O1 I2 I1 O R O1 I O2 O I1 I2 R O2 I O1 O I2I1 R O1 I O2 I2 O I1 R O2 I O1 I1 O I2 R O1 I O2 F I1 I2, O R O2 I O1 FI2, O I1 R O1 I O2 F I2, O I1 R O2 I O1 F I1 I2, O R O1 I O2 F I2 I1, OR O2 I O1 F I1, O I2 R O1 I O2 F I1, O I2 R O2 I O1 F I2 I1, O

FIG. 35 illustrates a perspective view of an embodiment of thetransmission 100 combined with a planetary gearset 505 in anoutput-coupled arrangement. In this output-coupled arrangement, theparallel path is eliminated and one or more sources of input torque arecoupled to the planetary gearset 505. The planetary gearset 505 then hasone or two outputs coupled with corresponding one or two of thecomponents of the CVT 100. For example, in one configuration, the ringgear 524 is rigidly attached to the case 40 (not shown), multiple planetgears 522 are operatively attached to the input disc 34 through theirplanet shafts 523, and the input is coupled to a planet carrier (notshown), which connects the planet shafts 523. The planet gears 522rotate the sun gear 520 in this arrangement, and the sun gear 520 isalso attached to a cage shaft 521, which rotates the cage 89 (notshown). The sun gear 520 rotates once each time the planet gears 522orbit the sun gear 520 and it is also rotated further by the planetgears 522 rotating about their respective axes 523. Therefore, the sungear 520 and the cage 89 (not shown) rotate faster than the planetcarrier (not shown) and the input disc 34.

Due to the fact that the cage 89 is rotating faster than the input disc34 in this configuration, the balls 1 rotate in the reverse direction ofthe input and the orientation of the variating components for the speedrange of the CVT 100 is reversed; the orientation for low speed of otherembodiments provides high speed here, and the orientation for high speedprovides low speed here. As the idler 18 (not shown) moves toward theinput side of the CVT 100, output speed can be decreased to zero and theoutput disc 101 will not rotate. In other words, this condition occurswhen a transmission is fully engaged with a rotating input but theoutput does not rotate. Such a condition can be obtained by adjustingthe tooth count of the planet gears 522 and sun gear 520. For example,if the sun gear 520 is twice the size of the planet gears 522, the sungear 520 and the cage 89 will rotate at twice the speed of the planetcarrier and the input disc 34. By increasing the cage 89 speed relativeto the input disc 34 speed, a range can be produced where the outputdisc 101 rotates in reverse at one end of the shift range of the CVT100, and where somewhere between this end and the midpoint of the shiftrange of the CVT 100, the output disc 101 speed is zero. The point inthe shift range of the CVT 100 where the output disc 101 speed is zerocan be plotted by dividing the speed of the sun gear 520 into the speedof the planet carrier, assuming that all other factors that determinethe shift range that provides a zero output speed are constant.

The following table, titled “Internally Coupled Planetary on Input,”shows most if not all of the combinations that can be achieved byvarying the embodiment illustrated in FIG. 35. For reference to thecomponents of the planetary gearset 505, “I1 refers to an outputcomponent that is coupled to a first CVT 100 input “I1,” “I2” refers toa second output component that is coupled to a second CVT 100 inputcomponent “I2,” and “F” refers to a component that is fixed for both theplanetary gearset 505 and the CVT 100. For the CVT 100, “R” refers to acomponent that is either free to rotate or is a second output of torque.In this table and the preceding table, only the planetary ring gear isshown as fixed and any of the planetary elements could be the fixedelement, which structure would result in more combinations. Suchadditional combinations are not shown herein to save space. Furthermore,in the table that follows, only one input from a prime mover (engine) isshown. This configuration has the capacity to accept two independentinputs thru the planetary, as in a parallel hybrid vehicle, but thesecombinations have not been illustrated separately in order to attempt toconserve space and it is understood that those in the art wouldapprehend these additional embodiments from the illustrated examples andthis statement. It should also be noted that any configuration from thefollowing table could be combined with any configuration from thepreceding table, either with single or dual cavity CVTs, to produce aset of configurations using two planetaries, one on the input and one onthe output. Internally Coupled Planetary on Input Planetary Gearset CVTI1 = Coupled to CVT I1 I1, I2 = Inputs from Planetary Gearset IT =Coupled to Input Torque from prime mover O = Output to vehicle orequipment F = Fixed to ground F = Fixed to ground I2 = Coupled to CVT I2R = Rolling free or 2nd output Input Output Disc Idler Cage DiscVariator Ring Carrier Sun (34) (18) (89) (101) Dual In/ I1 I2 IT I1 I2 FO Single Out I1 I2 IT I2 I1 F O IT I1 I2 I1 I2 F O IT I1 I2 I2 I1 F O I2IT I1 I1 I2 F O I1 IT I2 I2 I1 F O F I1 I2, IT I1 I2 F O F I2 I2, IT I2I1 F O F I2, IT I1 I1 I2 F O F I2, IT I2 I2 I1 F O F I2 I1, IT I1 I2 F OF I2 I1, IT I2 I1 F O F I1, IT I2 I1 I2 F O F I1, IT I2 I2 I1 F O DualIn/ I1 I2 IT I1 I2 R O Single Out I1 I2 IT I2 I1 R O IT I1 I2 I1 I2 R OIT I1 I2 I2 I1 R O I2 IT I1 I1 I2 R O I1 IT I2 I2 I1 R O F I1 I2, IT I1I2 R O F I2 I2, IT I2 I1 R O F I2, IT I1 I1 I2 R O F I2, IT I2 I2 I1 R OF I2 I1, IT I1 I2 R O F I2 I1, IT I2 I1 R O F I1, IT I2 I1 I2 R O F I1,IT I2 I2 I1 R O Dual In/ I1 I2 IT I1 I2 O F Single Out I1 I2 IT I2 I1 OF IT I1 I2 I1 I2 O F IT I1 I2 I2 I1 O F I2 IT I1 I1 I2 O F I1 IT I2 I2I1 O F F I1 I2, IT I1 I2 O F F I2 I2, IT I2 I1 O F F I2, IT I1 I1 I2 O FF I2, IT I2 I2 I1 O F F I2 I1, IT I1 I2 O F F I2 I1, IT I2 I1 O F F I1,IT I2 I1 I2 O F F I1, IT I2 I2 I1 O F Dual In/ I1 I2 IT I1 I2 O R SingleOut I1 I2 IT I2 I1 O R IT I1 I2 I1 I2 O R IT I1 I2 I2 I1 O R I2 IT I1 I1I2 O R I1 IT I2 I2 I1 O R F I1 I2, IT I1 I2 O R F I2 I2, IT I2 I1 O R FI2, IT I1 I1 I2 O R F I2, IT I2 I2 I1 O R F I2 I1, IT I1 I2 O R F I2 I1,IT I2 I1 O R F I1, IT I2 I1 I2 O R F I1, IT I2 I2 I1 O R Dual In/ I1 I2IT I1 O I2 F Single Out I1 I2 IT I2 O I1 F IT I1 I2 I1 O I2 F IT I1 I2I2 O I1 F I2 IT I1 I1 O I2 F I1 IT I2 I2 O I1 F F I1 I2, IT I1 O I2 F FI2 I2, IT I2 O I1 F F I2, IT I1 I1 O I2 F F I2, IT I2 I2 O I1 F F I2 I1,IT I1 O I2 F F I2 I1, IT I2 O I1 F F I1, IT I2 I1 O I2 F F I1, IT I2 I2O I1 F Dual In/ I1 I2 IT I1 O I2 R Single Out I1 I2 IT I2 O I1 R IT I1I2 I1 O I2 R IT I1 I2 I2 O I1 R I2 IT I1 I1 O I2 R I1 IT I2 I2 O I1 R FI1 I2, IT I1 O I2 R F I2 I2, IT I2 O I1 R F I2, IT I1 I1 O I2 R F I2, ITI2 I2 O I1 R F I2 I1, IT I1 O I2 R F I2 I1, IT I2 O I1 R F I1, IT I2 I1O I2 R F I1, IT I2 I2 O I1 R Dual In/ I1 I2 IT I1 F I2 O Single Out I1I2 IT I2 F I1 O IT I1 I2 I1 F I2 O IT I1 I2 I2 F I1 O I2 IT I1 I1 F I2 OI1 IT I2 I2 F I1 O F I1 I2, IT I1 F I2 O F I2 I2, IT I2 F I1 O F I2, ITI1 I1 F I2 O F I2, IT I2 I2 F I1 O F I2 I1, IT I1 F I2 O F I2 I1, IT I2F I1 O F I1, IT I2 I1 F I2 O F I1, IT I2 I2 F I1 O Dual In/ I1 I2 IT I1R I2 O Single Out I1 I2 IT I2 R I1 O IT I1 I2 I1 R I2 O IT I1 I2 I2 R I1O I2 IT I1 I1 R I2 O I1 IT I2 I2 R I1 O F I1 I2, IT I1 R I2 O F I2 I2,IT I2 R I1 O F I2, IT I1 I1 R I2 O F I2, IT I2 I2 R I1 O F I2 I1, IT I1R I2 O F I2 I1, IT I2 R I1 O F I1, IT I2 I1 R I2 O F I1, IT I2 I2 R I1 ODual In/ I1 I2 IT F I1 I2 O Single Out I1 I2 IT F I2 I1 O IT I1 I2 F I1I2 O IT I1 I2 F I2 I1 O I2 IT I1 F I1 I2 O I1 IT I2 F I2 I1 O F I1 I2,IT F I1 I2 O F I2 I2, IT F I2 I1 O F I2, IT I1 F I1 I2 O F I2, IT I2 FI2 I1 O F I2 I1, IT F I1 I2 O F I2 I1, IT F I2 I1 O F I1, IT I2 F I1 I2O F I1, IT I2 F I2 I1 O Dual In/ I1 I2 IT R I1 I2 O Single Out I1 I2 ITR I2 I1 O IT I1 I2 R I1 I2 O IT I1 I2 R I2 I1 O I2 IT I1 R I1 I2 O I1 ITI2 R I2 I1 O F I1 I2, IT R I1 I2 O F I2 I2, IT R I2 I1 O F I2, IT I1 RI1 I2 O F I2, IT I2 R I2 I1 O F I2 I1, IT R I1 I2 O F I2 I1, IT R I2 I1O F I1, IT I2 R I1 I2 O F I1, IT I2 R I2 I1 O

In the preceding tables, it is assumed that only one CVT 100 and onlyone planetary gearset 505 are being utilized. It is known in the art toutilize more planetary gearsets to develop additional combinations. Dueto the fact that the CVT 100 described in the tables can be implementedin a similar manner to a planetary gearset, it is easy for those ofskill in the art to combine the CVT 100 with a planetary gearset on bothits input and output ends in order to create substantially morecombinations, which combinations are known in the art and cannotreasonably be listed herein. However, such combinations are fully withinthe capabilities of those of skill in the art and are also to beconsidered as part of this description.

EXAMPLES

Each of these variations may have advantageous characteristics forparticular applications. The variations can be modified and controlledas necessary to achieve the goals for any particular application.Specific embodiments will now be described and illustrated that employsome of the variations described herein and/or listed in the abovetables. FIGS. 36 a, b, and c illustrate one embodiment of a transmission3600 that is a variation having one source of torque input and thatsupplies two sources of torque output. As before, only the significantdifferences between the embodiment illustrated in FIGS. 36 a, b and cand the previously illustrated and described embodiments will bedescribed. Furthermore, the components illustrated are being provided toillustrate to one of skill in the art how to provide power paths andtorque output sources that have not been previously illustrated. It isfully understood that many additional components can and will beutilized for operational embodiments, however for simplification of thedrawing, many such components have been omitted or are representedschematically as boxes.

Referring to FIG. 36 a, Torque is input through a drive shaft 3669 as inpreviously described embodiments. The drive shaft 3669 of thisembodiment is a hollow shaft having two ends and engaging on a first endwhatever prime mover is providing torque to the transmission 3600 andengaging at the second end a planet carrier 3630. The planet carrier3630 is a disc positioned coaxial with the longitudinal axis of thetransmission 3600 that interfaces at its center with the drive shaft3669 and extends radially to a radius near that of the inner side of thecase 3640 of the transmission 3600. In this embodiment, the case 3640 isstationary and is fixed to some supporting structure of the vehicle orequipment upon which it is utilized. A radial carrier bearing 3631 islocated between the inner surface of the case 3640 and the outer edge ofthe planet carrier 3630. The carrier bearing 3631 of some embodiments isa radial bearing that provides radial support to the planet carrier3630. In other embodiments, the carrier bearing 3631 is a compoundbearing providing both radial and axial support to the planet carrierpreventing cocking as well as radial or axial movement.

A plurality of planet shafts 3632 extend from the planet carrier 3630from a radial position between the center and the outer edge of theplanet carrier 3630. The planet shafts 3632 extend axially toward theoutput end of the transmission 3600 and are generally cylindrical shaftsthat connect the planet carrier 3630 to the input disc 3634 and eachform an axis about which a respective planet gear 3635 rotates. Theplanet shafts 3632 can be formed into the input side of the input disc3634 or the planet carrier 3630 or can be threaded into either the inputdisc 3634 or the planet carrier or can be attached by fasteners orotherwise. The planet gears 3635 are simple rotary gears that aresupported by and rotate about the planet shafts 3632 and manyembodiments utilize bearings between the planet gears 3635 and theplanet shafts 3632. They can have straight teeth or helical teeth,however where helical gears are used, thrust bearings are used to absorbthe axial thrust developed by the transmission of torque by the planetgears 3635.

Still referring to the embodiment illustrated in FIG. 36 a, the planetgears 3635 engage at two areas along their respective circumferences atany one time as they rotate about their respective axes. At a firstcircumferential position located farthest away from the longitudinalaxis of the transmission 36, each planet gear 3635 engages a ring gear3637. The ring gear 3637 is an internal gear formed on or attached tothe inner surface of the case 3640. In some embodiments, the ring gear3637 is a set of radial teeth formed on the inner surface of the ringgear 3637 and extending radially inward such that the planet gears 3635can engage with its teeth and ride along the inner surface of the ringgear 3637 as they orbit the longitudinal axis of the transmission 3600.At a circumferential point of the planet gears 3635 generally oppositethe radially outward most part, the ring gears 3635 engage a sun gear3620. The sun gear 3620 is a radial gear that is mounted coaxially aboutthe longitudinal axis of the transmission 3600 at the center of theplanet gears 3635 and engages all of the planet gears 3635. As theplanet carrier 3630 rotates the planet gears 3635 about the sun gear3620, the planet gears 3635 are rotated about their respective planetshafts 3632 by their engagement with the ring gear 3637 and thereforeboth orbit the sun gear 3620 and rotate on their own shafts as theyorbit. This results in a rotational energy that is transmitted to thesun gear 3620 that is at a greater speed than the speed input by thedrive shaft 3669.

In the embodiment illustrated in FIG. 36 a, the drive shaft 3669 alsodrives the input disc 3634 via the planet carrier 3630 and the planetshafts 3632. However, the planet gears 3635 also drive the sun gear 3620so that the power from the planet carrier is distributed to the inputdisc 3634 and the sun gear 3620. The sun gear 3620 is rigidly connectedto and rotates the cage 3689 of this embodiment. The cage 3689 issimilar to the embodiments described above, and therefore not all of thecomponents have been illustrated to simplify the drawing and improve theunderstanding of this description. The cage 3689, as in otherembodiments, positions the balls 3601 about the longitudinal axis of thetransmission 3600 and because the cage 3689 of this embodiment rotatesabout its axis, it causes the balls 3601 to orbit the longitudinal axisof the transmission 3600. The input disc 3634, which is similar to thosedescribed above provides an input torque to the balls 3601 in the samemanner as in previous embodiments. However the sun gear 3620 alsoprovides an input torque to the balls 3601 by rotating the cage 3689,which is added to the input from the input disc 3634. In thisembodiment, the output disc 3611 is rigidly fixed to the case 3640 anddoes not rotate about its axis. Therefore, the balls 3601 roll along thesurface of the output disc 3611 as they orbit the longitudinal axis ofthe transmission 3600 and rotate about their respective axes.

The balls 3601 cause the idler 3618 to rotate about its axis as in otherembodiments, however in this embodiment, the idler 3618 includes anidler shaft 3610 that extends out beyond the whole formed by the innerdiameter of the output disc 3611. The balls 3601 drive the idler 3618,which in turn drives the idler shaft 3610, which provides the firsttorque output from the transmission 3600. As illustrated in FIG. 36 b,the idler shaft 3610 can be of a cross-sectional shape that lends itselfto easier coupling with devices that would take power from the idlershaft 3610 and in some embodiments, as illustrated, the shape ishexagonal, although any such shape can be used. It is noted that due toaxial movement of the idler 3618 during shifting as described below, theidler shaft 3610 moves axially during shifting of the transmission 3600.This means that the couple between the idler shaft 3610 and the outputdevice (not shown) of this design allows for axial motion of the idlershaft 3618. This can be accomplished by allowing a slightly largeroutput device shaft such that the idler shaft 3610 is free to movewithin the output device, or by the use of a splined output idler shaft3610, such as by ball spline. Alternatively the idler 3618 can besplined to the idler shaft 3610 in order to maintain the axial positionof the idler shaft 3610.

Still referring to FIGS. 36 a and 36 b, the cage 3689 can provide anoutput power source as well. As illustrated, the cage 3689 can beconnected on its inner diameter on the output side to a cage shaft 3690.In the illustrated embodiment, the cage shaft 3690 is formed at its endinto an output gear or spline to engage and supply power as a secondoutput source.

As illustrated in FIG. 36 a, various bearings can be implemented tomaintain the axial and radial position of various components in thetransmission 3600. The cage 3689 can be supported in its place by cageoutput bearings 3691, which are either radial bearings to provide radialsupport or are preferably combination bearings to maintain both axialand radial position of the cage with respect to the case 3640. The cageoutput bearings 3691 are assisted by cage input bearings 3692 which arealso radial or preferably combination radial-thrust bearings andposition the cage 3689 relative to the input disc 3634. In embodimentsutilizing an axial force generator where the input disc 3634 is subjectto slight axial movement or deformation, the cage input bearings 3692are designed to allow for such movement by any mechanism known in theindustry. One embodiment utilizes an outer bearing race that is splinedto the inner diameter of the input disc 3634, by a ball spline forexample, in order that the input disc 3634 can move axially slightlyrelative to the outer race of the cage input bearing 3692.

The shifting mechanism of the embodiment illustrated in FIG. 36 a isslightly varied from the embodiments illustrated to allow for the outputtorque supplied by the idler 3618. In this embodiment, the idler 3618initiates the shifting by being moved axially upon actuation by theshift rod 3671 and in turn moves the shift guides 3613 axially causingthe shifting mechanism to change the axes of the balls 3601 as describedabove. The shift rod 3671 does not thread into the idler 3618 in thisembodiment, however and only contacts the idler 3618 via idler inputbearings 3674 and idler output bearings 3673. The idler input and outputbearings 3674, 3673, respectively, are combination thrust and radialbearings that position the idler 3618 both radially and axially alongthe longitudinal axis of the transmission 3600.

When the shift rod 3671 is moved axially toward the output end, theinput idler bearing 3674 apply axial force to the idler, thereby movingthe idler axially to the output end and initiating a change in thetransmission ratio. The shift rod 3671 of the illustrated embodimentextends beyond the idler 3618 through an inner diameter formed in thecenter of the sun gear 3620 and into the second end of the drive shaft3669 where it is held in radial alignment within the drive shaft 3669 byan idler end bearing 3675. The shift rod 3671 moves axially within thedrive shaft 3669 however and therefore the idler end bearing 3675 ofmany embodiments allows for this motion. As described before, many suchembodiments utilize a splined outer race that engages a mating splineformed on the inner surface of the drive shaft 3669. This splined raceallows the race to slide along the inner surface of the drive shaft 3669as the shift rod 3671 is moved axially back and forth and still providesthe radial support used to assist in radially aligning the shift rod3671. The inner bore of the sun gear 3620 can also be supported radiallywith respect to the shift rod 3671 by a bearing (not illustrated)located between the shift rod 3671 and the sun gear 3620. Again eitherthe inner or outer race could be splined to allow for the axial motionof the shift rod 3671.

When the idler 3618 of the embodiment illustrated in FIG. 36 a is movedaxially to shift the transmission 3600, the idler 3618 moves the shiftguides 3613. In the illustrated embodiment, the shift guides 3613 areannular rings coaxially mounted about each end of the idler 3618. Theillustrated shift guides 3613 are each held in radial and axial positionby an inner shift guide bearing 3617 and an outer shift guide bearing3672. The inner and outer shift guide bearings of this embodiment arecombination bearings providing both axial and radial support to theshift guides 3613 in order to maintain the axial and radial alignment ofthe shift guides 3613 in relation to the idler 3618. Each of the shiftguides 3613 can have a tubular sleeve (not shown) that extends away fromthe idler 3618 so that the shift guide bearings 3617 and 3672 can befurther apart to provide additional support to the shift guides 3613, asneeded. The shift rod 3671 can be moved axially by any known mechanismfor causing axial motion such as an acme threaded end acting as a leadscrew or a hydraulically actuated piston or other know mechanisms.

Referring to FIGS. 36 a and b and mainly to FIG. 36 c, the paths ofpower through the transmission 3600 follow to parallel and coaxialpaths. Initially, power enters the transmission 3600 via the drive shaft3669. The power is then split and transmitted through the planet carrier3630 both to the input disc 3634 and to the sun gear 3620 via the planetgears 3635. The latter power path is then transmitted from the sun gear3620 to the cage 3689 and out of the transmission 3600 via the cageshaft 3689. This power path provides a fixed transmission ratio from thedrive shaft based upon the dimensions of the sun gear 3620 and theplanet gears 3635. The second power path is from the planet carrier 3630through the planet shafts 3632 and to the input disc 3634. This powerpath continues from the input disc 3634 to the balls 3601 and from theballs 3601 to the idler shaft 3618 and out of the transmission 3600through the idler shaft 3610. This unique arrangement allows the twopower paths to be transmitted through the transmission 3600 not only inparallel paths but through coaxial paths. This type of powertransmission allows for a smaller cross-sectional size for the sametorque transmission and leads to significant size and weight reductionsinto a much simpler design compared to other IVTs.

The embodiment illustrated in FIGS. 36 a, b and c, illustrates to one ofskill in the art how the idler 3618 can be used as a power output aslisted in the tables above and how to combine the planetary gear setwith the CVT as described above. It is expected that variations of thisdesign can be utilized while achieving the various combinationsdescribed, and such alternate designs cannot all be illustrated hereindue to the overwhelming number of combinations listed that areavailable. It is also understood that the axial force generatorsprovided herein can also be utilized with this embodiment, but forsimplification these devices are not illustrated. For embodimentsutilizing one of the axial force generators described herein, oranother, it is expected that the components of the axial force generatorcan be implemented between where the planet shafts 3632 connect to theinput disc 3634, although other arrangements can be employed as well. Insuch embodiments, the parallel path described in FIGS. 32 and 33 ismoved in to be coaxial with the axis of the transmission 3600 allowingfor a much smaller transmission 3600 for the same torque transmissionand thereby leading to reduced weight and space of such embodiments.FIGS. 36 a, b and c illustrate one combination in order to show howrotational power might be taken from the various components of thetransmission in various embodiments. Obviously, those of skill in theart will easily understand how other configurations provided herein canbe achieved by varying the connections, and it would be unnecessarilyburdensome and voluminous to illustrate all or even more combinationsfor the simple purpose of illustrating the combinations described. Theembodiments shown in FIGS. 35 and 36 a can therefore be modified asnecessary to produce any of the variations listed above or below withoutthe need for a separate non-coaxial parallel power path.

Referring now to FIG. 37 a, an alternative embodiment of a transmission3700 is illustrated. In this embodiment, the output disc 3711 is formedas part of the case of previous embodiments to form a rotating hub shell3740. Such an embodiment is suited well for applications such asmotorcycles or a bicycle. As mentioned before, only the substantialdifferences between this embodiment and the previously describedembodiments will be described in order to reduce the size of thisdescription. In this embodiment, the input torque is supplied to aninput wheel 3730, which can be a pulley for a belt or a sprocket for achain or some similar device. The input wheel 3770 is then attached tothe outside of a hollow drive shaft 3769 by press fitting or splining orsome other suitable method of maintaining angular alignment of the tworotary components. The drive shaft 3769 passes through a removable endof the hub shell 3740 called the end cap 3741. The end cap is generallyan annularly shaped disc having a bore through its center to allowpassage of the drive shaft 3769 into the inside of the transmission 3700and having an outer diameter that mates with the inner diameter of thehub shell 3740. The end cap 3741 can be fastened to the end cap 3740 orit can be threaded into the hub shell to encapsulate the innercomponents of the transmission 3700. The end cap 3741 of the illustratedembodiment has a bearing surface and corresponding bearing on the insideof its outer diameter for positioning and supporting the axial forcegenerator 3760 and has a bearing surface and corresponding bearing atits inner diameter that provides support between the end cap 3741 andthe drive shaft 3769.

The drive shaft 3769 fits over and rotates about an input axle 3751,which is a hollow tube that is anchored to the vehicle frame 3715 by aframe nut 3752 and that provides support for the transmission 3700. Theinput axle 3751 contains the shift rod 3771, which is similar to theshift rods described in previous embodiments, such as that illustratedin FIG. 1. The shift rod 3771 of this embodiment is actuated by a shiftcap 3743 threaded over the end of the input axle 3751 that extendsbeyond the vehicle frame 3715. The shift cap 3743 is a tubular cap witha set of internal threads formed on its inner surface that mate with acomplimentary set of external threads formed on the outer surface of theinput axle 3751. The end of the shift rod 3771 extends through a holeformed in the input end of the shift cap 3743 and is itself threadedallowing the shift cap 3743 to be fastened to the shift rod 3771. Byrotating the shift rod 3771 its threads, which may be acme threads orany other threads, cause it to move axially and because the shift rod3771 is fastened to the shift cap 3743, the shift rod 3771 is movedaxially as well, actuating the movement of the shift guides 3713 and theidler 3718, thereby shifting the transmission 3700.

Still referring to the embodiment illustrated in FIG. 37 a, the driveshaft 3769 rides on and is supported by the input axle 3751 and one ormore shaft support bearings 3772, which can be needle bearings or otherradial support bearings. The drive shaft 3769 provides torque to anaxial force generator 3760 as in previous embodiments. Any of the axialforce generators described herein can be used with this transmission3700, and this embodiment utilizes a screw 3735 that is driven by thedrive shaft 3769 by splining or other suitable mechanism thatdistributes torque to the drive disc 3734 and to a bearing disc 3760, asin the previous embodiments. In this embodiment, a drive seal 3722 isprovided between the inner diameter of the input wheel 3770 and theouter diameter of the input axle 3751 beyond the end of the drive shaft3769 in order to limit the amount of foreign material that is admittedto the inside of the transmission 3700. Another seal (not shown) can beused between the case cap 3742 and the input wheel to limit foreignparticle infiltration from between the end cap 3741 and the drive shaft3769. The drive seal 3722 can be an o-ring seal, a lip seal or any othersuitable seal. The illustrated embodiment also utilizes a similar cage3789 as previously described embodiments however, the illustratedtransmission 3700 utilizes axle bearings 3799 to support the balls 1 ontheir axles 3703. The axle bearings 3799 can be needle bearings or othersuitable bearings and reduce the friction between the balls and theiraxles 3703. Any of the various embodiments of balls and ball axlesdescribed herein or known to those of skill in the art can be used toreduce the friction that is developed.

Still referring to the embodiment illustrated in FIG. 37 a, the cage3789 and the shift rod 3771 are supported on the output side by anoutput axle 3753. The output axle 3753 is a somewhat tubular supportmember located in a bore formed in the output end of the hub shell 3740and between the cage 3789 and the output side vehicle frame 3715. Theoutput axle 3753 has a bearing race and bearing formed between its outerdiameter and the inner diameter of the hub shell 3740 to allow forrelative rotation of the two components as the output axle 3753 providessupport to the output side of the transmission 3700. The output shaft isclamped to the vehicle frame 3715 by an output support nut 3754.

As is illustrated in FIG. 37 a, this transmission 3700 is shifted byapplying tension to the shifting cord 3755 that is wrapped around andwhich applies rotational force to the shift cap 3743. The shift cord3755 is a tether capable of applying a tension force and is actuated bya shifter (not shown) used by the operator to shift the transmission3700. In some embodiments the shift cord 3755 is a guide wire capable ofboth pulling and pushing so that only one coaxial guide line (not shown)needs to be run to the shifter from the transmission 3700. The shiftingcord 3755 is conducted by housing stops 3716 to and from the shift capfrom the shifter used by the operator. The housing stops 3716 areextensions from the vehicle frame 3715 that guide the shifting cord 3755to the shift cap 3743. In the illustrated embodiment, the stop guides3716 are somewhat cylindrically shaped extensions having a slot formedalong their length through which the shifting cord 3755 passes and isguided. In other respects, the transmission 3700 illustrated in FIG. 37a is similar to other embodiments illustrated herein.

Another embodiment that is similar to the one illustrated in FIG. 37 ais illustrated in FIG. 37 b. In this embodiment, the output disc 3711 isalso fixed to the case 3740, however, the case 3740 is fixed and doesnot rotate. In this embodiment, however, similar to the embodimentillustrated in FIG. 36 a, the cage 3789 is free to rotate relative tothe output disc 3711 and the case 3740. This means that the output isagain through the idler 3718. In this embodiment the idler 3718 isattached to a moveable output shaft 3753 similar to that described inthe embodiment of FIG. 36 a. The output shaft 3753 terminates at the farend on the output side in an output spline 3754, which allows couplingof the moveable output shaft 3753 to whatever device is being suppliedwith torque by the transmission 3700. In this embodiment, torque issupplied to the transmission 3700 via the input shaft 3772 by a chainand sprocket (not shown), by an input gear (not shown) or by other knowncoupling means. The torque then passes through to the input disc 3734 asdescribed in the preceding embodiment. However, as described, withreference to FIG. 37 a, the balls 3701 ride along the surface of theoutput disc 3711 and transfer torque to the idler 3718.

As with the embodiment illustrated in FIG. 36 a, by supplying the torqueoutput via the idler 3718, the shift guides 3713 of this embodiment aresupported by bearings 3717 on the outer surface of the output shaft3753. This transmission 3700 is shifted by moving the shift rod 3771axially and is actuated by an actuator 3743. The actuator can be theshift cap of FIG. 37 a, or a wheel or gear controlled by an actuatingmotor or manually, or the actuator 3743 can be any other mechanism foraxially positioning the shift rod 3771, such as one or more hydraulicpistons. In some embodiments, the axial force generator 3960 and theshifting mechanism illustrated below in FIG. 39 a is utilized. Throughthis embodiment, a very high transmission ratio can be achieved at avery high efficiency and with very little frictional losses whencompared with other transmission types.

FIG. 38 illustrates an alternative embodiment of a ball axle 3803 thatcan be used with many of the transmission described herein. In thisembodiment, oil is pumped into the bore in the ball 1 by threads 3810formed in the outer diameter of the ball axle 3803. A layer of oil thatis adhered to the surface of the ball 1 in the vicinity of the bore, itis drawn about the axle 3803 as the ball 1 rotates and travels about theaxle 3803 at the same speed as the surface to which it is adhered; itadditionally draws adjacent layers of oil that are bound at everdecreasing binding strength, depending on their respective distancesfrom the surface layer, by the same attractive forces creating theviscosity of the oil. As these layers of oil are drawn about the axle,the leading edge of any particular volume of oil in a layer is shearedby the surface of a set of threads 3810 formed on the outer surface ofthe axle 3803. The threads 3810 can be acme threads or any other type ofthreads suitable for the pumping action described herein. As each volumeof oil is sheared from the adjacent layer that is outside the radius ofthe threads 3810, it is displaced by a similar layer that is shearedsubsequently by the same action. Because the threads 3810 are shaped sothey lead into the bore of the ball 1, the volumes of oil that aresheared moved inside the ball 1 as they are continually displaced byfurther shearing action occurring behind them. As this continues, theoil is forced inside of the bore of the ball 1 by its ownself-attractive forces and creates a sort of pumping action. This“pumping” action is therefore proportional to the viscosity of the oil.In order to facilitate this pumping effect, in many embodiments,lubricants are selected for use that act as Newtonian fluids in theshear rates experienced over the range of spin rates experienced by theballs 1 of any particular embodiment.

Still referring to FIG. 38, the threads 3810 begin at a point along theaxis of the ball axle 3803 that is slightly outside the edge of the ball1 in order to create the displacing shearing action that causes the oilto flow into the ball 1. The distance outside the ball 1 that thethreads 3810 extends can be between 0.5 thousandths of an inch and 2inches, while in other embodiments the distance can be from 10thousandths of an inch to one inch, or more or less depending uponmanufacturing costs and other considerations. The threads 3810 of theillustrated embodiment extend into the bore of the ball 1 and stopsomewhere inside the ball in a reservoir 3820 formed by a longitudinallength of the ball axle 3803 that is of a smaller diameter than the restof the ball axle 3803. This reservoir 3820 ends inside of the ball 1 ata reservoir end 3830 where the outer diameter of the ball axle againincreases to near the inner diameter of the ball 1 so that the oil isforced to leak out of the ball 1 from the small gap between the ballaxle 3803 and the inner surface of the ball 1 resulting in a highpressure oil supply for forming a lubricating film between the twocomponents. In some embodiments, a reservoir 3820 is not present and thethreads 3810 simply end in the vicinity of the middle of the bore.

An equilibrium can be developed between the amount of oil that leaks outand the amount that is pumped in to maintain a lubricating pressure inthe bore of the ball 1 by controlling the size of the gap between theball axle 3803 and the inner surface of the ball 1. This equilibrium isdependent upon the viscosity of the oil, the size of the gap and therotation rate of the ball 1. Although the reservoir end 3830 isillustrated as being located near the middle of the ball 1, this is onlyfor illustrative purposes and the reservoir 3820 can end closer to theother end of the ball 1 or nearer the threads 3810 depending upon theapplication. In other similar embodiments, this same orientation isformed by threads formed on the interior of the bore through the balls1, similar to that illustrated in FIG. 23, except that threads 3810 areformed as described in the present embodiment that end in a reservoir3820 formed near the middle of the ball 1 and ball axle 3803.

Referring now to FIGS. 39 a, b and c, another alternative axial forcegenerator 3960 is illustrated. In this embodiment the screw 3935 islocated in the inner bore of the bearing disc (not shown) instead of theinput disc 3934. In this embodiment, the screw 3935 is driven directlyby the drive shaft (not shown) via splines 3975, which mate withmatching splines from the drive shaft. The screw 3935 then distributestorque to the input disc 3934 via central screw ramps 3998 and centraldisc ramps 3999 and to the bearing disc via its threads 3976 and acorresponding set of internal threads (not shown) formed on the innersurface of the bore of the bearing disc. As the screw 3935 is rotated bythe drive shaft, a set of central screw ramps 3998 that are formed onthe output end of the screw 3935 are rotated and engage and rotate acomplimentary set of central disc ramps 3999. The central disc ramps3999 are formed on a thrust washer surface formed on the input side ofthe input disc 3934 near its inner diameter, and as they are rotated bythe central screw ramps 3998, the central disc ramps 3999 begin to applytorque and axial force to the input disc 3934 from the reaction of theangled surfaces of the central ramps 3998, 3999. Additionally, therotation of the screw 3935 causes its threads 3976 to engage with thethreads of the bearing disc to begin to rotate the bearing disc.

Referring now to FIG. 39 a in the illustrated embodiment, the axialforce generator 3960 is directly affected by the position of the idler3918. In this embodiment, the idler assembly has a tubular extensioncalled a pulley stand 3930 that extends from the input side thrust guide3713 and that ends near the input disc 3934 in an annular extensionspreading radially outward. A linkage assembly made up of a fixed link3916, a first link pin 3917, a short link 3912, a cam link 3914, a camlink pin 3915 and a stationary cam pin 3923 extends axially toward thescrew 3935 from the pulley stand 3930 and positions the screw 3935axially depending on the transmission ratio. The links 3916, 3912 and3914 are generally elongated struts. The fixed link 3916 extends fromthe input end of the pulley stand 3930 toward the screw 3935 and isconnected to the intermediate short link 3912 by the first link pin3917. The first link pin 3917 forms a floating pin joint between thefixed link 3916 and the short link 3912 such that the short link 3912can rotate about the first link pin 3917 as the two links 3916, 3912move axially during shifting. The short link 3912 is then connected atits other end to the cam link 3914 by a cam link pin 3915 and therebyforms a floating pin joint. The cam link 3914 is fixed axially by astationary cam pin 3923 that is fixed to the axle 3971 or anotherstationary component and forms a pin joint about which the cam link 3914rotates as the idler 3918 moves axially.

In the following description, for simplification of the drawing, thebearing disc 60, ramp bearings 62, perimeter ramps 61 and input discramps 64 of FIG. 1 are not separately illustrated, but similarcomponents can be utilized to fulfill similar functions in the presentembodiment. When the axial force generator 3960 illustrated in FIGS. 39a, b and c is in a high transmission ratio, the idler 3918 is located atan axial position at its far input side and therefore the fixed link3916 is also located its farthest axial point toward the input side. Thefirst link pin 3917, the short link 3912 and the second link pin 3921are all located towards the input side and therefore the cam link 3914is oriented about the stationary cam pin 3923 such that its cam surface(not separately illustrated) is rotated away from the screw 3935. Thecam link 3914 applies cam force to the screw 3935 when it is rotatedabout its fixed stationary cam pin 3923 axis to force the screw towardthe output side when in low transmission ratios. However in lowtransmission ratios, as illustrated, the cam surface of the cam link3914 is rotated away from the screw 3935. This allows the screw 3935 tosettle at its farthest point towards the output side and results in thebearing disc rotating counter-clockwise, looking from the input sidetowards the output side, about the screw 3935 in order to maintainengagement with the screw threads 3976. As this occurs the bearing rampsare rotated counter-clockwise allowing the disc bearings (notillustrated here but similar to those previously described with respectto FIG. 1) to roll to a point between the bearing disc ramps and theramps of the input disc 3934 where the bearings provide little or noaxial force.

Meanwhile, due to the extreme position of the screw 3935 to the left asviewed in FIG. 39 a, the central screw ramps 3998 are engaged with thecentral disc ramps 3999 fully such that the input disc 3934 is rotatedclockwise slightly to allow the axial position of the screw 3935 in itsfarthest output side position. The rotation of the input disc 3934 inthis manner means that the input disc ramps have rotated in an oppositedirection of the bearing disc ramps thereby amplifying the effect ofunloading the perimeter ramps and bearings. In such a situation, themajority or all of the axial force is being applied by the central ramps3998, 3999 and little if any axial force is generated by the perimeterramps.

As the idler 3918 moves toward the output side to shift to a lowertransmission ratio, the linkage assembly becomes extended as the fixedlink 3916 moves axially away from the screw 3935, and the cam link 3914is rotated about the stationary cam pin 3923. As the cam link 3914 isrotated about the cam link pin 3923, the axial motion of the fixed link3916 acts upon one end of the can link 3914, while the other end movestoward the screw 3935, thereby reversing the direction of the axialforce applied by the fixed link 3916. By adjusting the lengths of wherethe various connections are made to the cam link 3914, the axial forceapplied by the fixed link 3916 can be diminished or magnified by leveraction. The cam end of the cam link 3914 applies an axial force to athrust washer 3924 on the output side of the screw 3935. The thrustwasher 3924 engages a screw thrust bearing 3925 and a bearing race 3926to supply the resultant axial force to the screw 3935. In response, thescrew 3935 moves axially toward the input side and its threads 3976rotate the bearing disc clockwise, looking from input side to outputside, causing the perimeter ramps to rotate so that the ramp bearingsare moved along the perimeter ramps to a position where they begin todevelop axial force. At the same time, due to the axial movement of thescrew 3935 toward the input side, the central screw ramps 3998 aredisengaged from the central disc ramps 3999 and the input disc 3934rotates, relative to the screw 3935, counter-clockwise, again aiding themovement of the perimeter ramp bearings to a position to generate axialforce. Through this lever action of the linkage assembly, the axialforce generator 3960 of this embodiment efficiently distributes theaxial force and torque between the central ramps 3998, 3999 and theperimeter ramps.

Also illustrated in FIG. 39 a is an alternative leg assembly to that ofFIG. 5 that allows for a reduced overall size of the transmission. Inthe illustrated embodiment, the rollers 3904 are positioned radiallyinward on the legs 3902 as compared to the legs 2 of FIG. 5.Additionally, the input disc 34 and output disc (not shown) contact theballs 1 at a point closer to their axes which reduces the load on theidler 18 and enables the transmission to carry more torque. With thesetwo modifications, the input disc 34 and output disc (not shown) of thisembodiment can be reduced in total diameter to a diameter substantiallythe same as the farthest opposing points on two diametrically opposingballs 3901 of this embodiment as illustrated by the line “O.D.”

Another feature of the embodiment illustrated in FIG. 39 a is a modifiedshifting assembly. The rollers 3904 of this embodiment are formed aspulleys each with a concave radius 3905 at its outer edge instead of aconvex radius. This allows the rollers 3904 to fulfill their function ofaligning the ball axles 3903 but also allows them to act as pulleys tochange the axes of the ball axles 3903 and the balls 3901 in order toshift the transmission. The flexible cables 155 described with respectto FIGS. 1 and 6, or similar shifting cables can be wrapped around therollers 3904 of one side so that when a tension is applied, thoserollers 3904 come closer together, thereby shifting the transmission.The shifting cables (not illustrated in FIG. 39) can be guided throughthe cage (item 89 of FIG. 1) to the rollers 3904 by guide rollers 3951,which in the illustrated embodiment are also pulleys mounted on guideshafts 3952 to the output end of the pulley stand 3930.

In some embodiments, the guide rollers 3951 and the guide shafts 3952are designed to allow the axis of the guide rollers 3951 to pivot inorder to maintain a pulley-type alignment with the rollers 3904 as theball axles 3903 change their angles with respect to the axis of thetransmission. In some embodiments, this can be accomplished by mountingthe guide shafts 3952 to the pulley stand 3930 with pivot joints ortrunnions, or any other known method. In this embodiment, one shiftcable can act on one set of rollers 3904 on either the input side or theoutput side of the balls 3901 and a spring (not shown) biases the ballaxles 3903 to shift in the other direction. In other embodiments, twoshifting cables are used with one on one side that draws the rollers3904 on its side radially inward and another cable on the opposite endof the balls 3901 that draws the rollers 3904 on its respective sideradially inward shifting the transmission thusly. In such an embodimenta second pulley stand 3930 or other suitable structure is formed on theoutput end of the shift guides 3913 and a corresponding set of guideshafts 3925 and guide rollers 3951 is mounted on that second pulleystand 3930. The cables (not shown) of such embodiments pass throughholes or slots (not shown) formed in the axle 3971 and out of thetransmission via the axle 3971. The cables can pass out of either orboth of the ends of the axle 3971 or they can pass out of additionalholes formed through the axle 3971 axially beyond either or both theinput disc (not shown) and the output disc (also not shown), or the hub(not shown) it the output disc is a rotating hub. The holes and or slotsthrough which the cables pass are designed to maximize the life of thecable material through the use of radiused edges and pulleys and suchdevices are used in various locations of the axle and transmission forconveyance of the cable.

Referring to FIGS. 39 a, 40 a and 40 b, one embodiment of a linkageassembly 4000 of the axial force generator 3960 of FIG. 39 a isillustrated. The illustrated linkage assembly 4000 is also made up of afixed link 3916, a first link pin 3917, a short link 3912, a second linkpin 3921, a cam link 3914, a cam link pin 3915 and a stationary cam pin.The fixed link 3916 of this embodiment is an elongated strut having afirst end that is rigidly attached to the pulley stand 3930 of FIG. 39a, and a second end facing away from the first end that has a pin jointhole formed through it. The fixed link 3916 is generally parallel alongside the axle 3971. The first link pin 3917 is placed within the hole inthe second end of the fixed link 3916 joining the second end of thefixed link 3916 with a first end of the short link 3912, which has acorresponding pin joint hole formed therein. The sort link 3912 is alsoa strut having two ends, however it has holes formed in both its firstand second end. A cam link pin 3915 is placed within the hole in thesecond end of the short link 3912 and joins the second end of the shortlink 3912 with the first end of the cam link 3914 via a pin joint holeformed in the cam link 3914. The cam link 3914 has two ends, a first endand an opposite cam end that has a cam surface 4020 formed upon itsouter edge. The cam link 3914 also has a second pin joint hole formedmidway between its first end and the cam end through which thestationary cam pin 3923. The stationary cam pin 3923 is fixed to astationary part of the transmission such as the axle 3971 so that itforms an axis about which the cam link 3914 rotates.

FIG. 40 a illustrates the linkage assembly 4000 in its contracted statecorresponding to a very high transmission ratio, where the fixed linkhas moved all the way toward the input end of the transmission asdescribed above for FIG. 39 a. FIG. 40 b illustrates the linkageassembly 4000 in an extended state corresponding to a low transmissionratio. As was described above, the cam link 3914 applies an axial forceto the screw 3935 in order to shift the generation of the axial forcefrom the central ramps 3998, 3999 to the perimeter ramps as thetransmission is shifted from high to low. Additionally, when thetransmission is shifted from low to high, the cam link 3914 reduces theamount of axial force that is applied to the screw 3935 allowing thescrew 3935 to move axially toward the output end and thereby shift theaxial force generation from the perimeter ramps in to the central ramps3998, 3999.

As is illustrated in FIGS. 40 a and b, the cam surface 4020 of the camlink 3914 can be designed to provide a great variety of loading andunloading profiles. In fact, in this embodiment, a second cam surface4010 is provided on the first end of the cam link 3914. As illustratedin FIG. 40 a, at a very high transmission ratio, the cam surface 4020 isfully unloaded applying a minimal amount of, if any at all, axial forceto the screw 3935. However, in some embodiments, a higher amount ofaxial force may need to be applied at various speed ratios, and in thiscase, at the highest transmission ratio the second cam surface 4010increases the axial force to the screw thereby transferring some axialforce generation back to the perimeter discs to increase the amount ofaxial force that may be needed at that high transmission ratio. This ismerely an example of the variations that can be included to vary thecontrol of the generation of axial force by the axial force generator4060 depending on the desired torque-speed profile of a particularapplication.

The embodiments described herein are examples provided to meet thedescriptive requirements of the law and to provide examples. Theseexamples are only embodiments that may be employed by any party and theyare not intended to be limiting in any manner. Therefore, the inventionis defined by the claims that follow and not by any of the examples orterms used herein,

1. A bicycle comprising: a bicycle frame; a plurality of bicycle wheelsoperationally coupled to the bicycle frame; a power train operationallycoupled to the wheels, the power train comprising: a planetary gearsetmounted on a shaft, the shaft defining a longitudinal axis; a variatorcomprising: a plurality of balls, each ball having a tiltable axis aboutwhich it rotates; a first disc in contact with the balls; a second discin contact with the balls; and an idler in contact with the balls;wherein the variator is mounted coaxially with the planetary gearsetalong the longitudinal axis; and wherein the planetary gearset isoperationally coupled to the variator.
 2. The bicycle of claim 1,further comprising a power source operationally coupled to the powertrain.
 3. The bicycle of claim 2, wherein the power source comprises atleast one of an electric motor and an internal combustion engine.
 4. Thebicycle of claim 1, wherein the planetary gearset comprises: a ring gearmounted coaxially about the longitudinal axis; a plurality of planetgears in engagement with the ring gear, each planet gear having a planetshaft about which the planet gear rotates; a sun gear mounted coaxiallyabout the longitudinal axis and in engagement with the planet gears; anda planet carrier mounted coaxially about the longitudinal axis andadapted to support and position the planet shafts.
 5. The bicycle ofclaim 4, wherein the planet carrier is operationally coupled to a driveshaft.
 6. The bicycle of claim 5, wherein the drive shaft receives aninput torque from a sprocket, gear, or crank that is operationallycoupled to a prime mover.
 7. The bicycle of claim 1, further comprisingshift guide sleeves operationally coupled to the idler.
 8. The bicycleof claim 7, further comprising a shifting mechanism that shifts thetransmission by axially moving the guide sleeves.
 9. The bicycle claim4, further comprising a cage adapted to align the tiltable axes of theballs and to maintain the angular and radial positions of the balls. 10.The bicycle of claim 9, wherein the cage is nonrotatable.
 11. Thebicycle of claim 9, wherein the cage is adapted to rotate about thelongitudinal axis.
 12. The bicycle of claim 9, wherein the idler or thecage provides a torque input to the planetary gearset.
 13. The bicycleof claim 9, wherein the planet carrier receives an input torque and iscoupled to the first disc, wherein the sun gear is coupled to the cage,wherein the ring gear is fixed and does not rotate, and wherein thevariator supplies an output torque via the second disc.
 14. The bicycleof claim 4, further comprising a force generator adapted to generate anaxial force that increases the traction between the first and seconddiscs, the balls, and the idler.
 15. The bicycle of claim 14, whereinthe force generator produces an amount of axial force that is a functionof a transmission ratio of the variator.
 16. The bicycle of claim 1,further comprising a shaft operationally connected to the idler andadapted to receive a torque from the idler and to transmit the torqueout of the variator.
 17. The bicycle of claim 1, further comprising aforce generator that comprises: a third disc having a threaded bore; afirst set of ramps operably coupled to the third disc near an outerdiameter of the third disc; a plurality of bearings adapted to engagethe first set of ramps; and a second set of ramps, the second set oframps operably coupled to the first disc and adapted to engage thebearings.
 18. The bicycle of claim 17, the force generator furthercomprising: a rotatable screw adapted to engage the threaded bore; athird set of ramps, the third set of ramps operably coupled to thescrew; and a fourth set of ramps, the fourth set of ramps operablycoupled to the first disc and adapted to engage the third set of ramps.19. The bicycle of claim 18, the force generator further comprising: athrust bearing contacting the screw; and a linkage assembly adapted toapply a force to the thrust bearing tending to move the screw away fromthe first disc in response to a shift in the transmission ratio.
 20. Thebicycle of claim 19, wherein in response to a shift in the transmissionratio the linkage assembly varies the axial force applied to the thrustbearing as a function of the transmission ratio.